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ASHRAE Research Project ReportRP-394 A Study to Determine Methods for Designing Radiant Heating and Cooling Systems Approval: May 1987 Contractor: University of Missouri-Rolla Rolla, MO 65401 Principal Investigator: Authors: Ronald Howell N/A Author Affiliations, Sponsoring Committee: TC 6.5, Radiant Space Heating and Cooling Co-Sponsoring Committee: Co-Sponsoring Organizations: N/A N/A ©2012 ASHRAE www.ashrae.org. This material may not be copied nor distributed in either paper or digital form without ASHRAE’s permission. Requests for this report should be directed to the ASHRAE Manager of Research and Technical Services. FINAL REPORT ASHRAE RP - 394 A STUDY TO DETERMINE METHODS FOR DESIGNING RADIANT HEATING AND COOLING SYSTEMS Ronald H. Howell Department of Mechanical & Aerospace Engineering University of Missouri-Rolla Rolla, MO 65401 May 1987 I ^«SWC>WS0C»ETV < S^" GH T TABLE OF CONTENTS SUMMARY 1 1.0 - OBJECTIVES AND SCOPE 3 2.0 - INTRODUCTION 4 3.0 - DEFINITIONS AND TERMINOLOGY 3.1 - Radiation, Convection and Conduction 3.2 - Infrared Ranges 3.3 - Low, Medium, and High Temperature Radiant Sources 3.4 - Mean Radiant Temperature, MRT 3.5 - Radiant Temperature Asymmetry 3.6 - Operative Temperature 3.7 - Effective Radiant Flux 3.8 - Average Unheated Surface Temperature 3.9 - Comfort Conditions 3.10 - Design Heat Loss Values 3.10.1 3.10.2 3.10.3 3.10.4 - ASHRAE Standard Heat Loss (HLD) Actual Design Heat Loss (HLA) Conduction Design Heat Loss (HLC) Conduction Design Heat Loss with Room Air Temperature Gradient (HLCG) 4.0 - BACKGROUND 16 16 17 19 20 20 21 22 22 22 23 23 24 24 24 25 4.1 - Relationship Between Radiant Heating/Cooling and Comfort Conditions. 25 4.1.1 - Fanger's Comfort Equation 25 4.1.2 - Changes in Air Temperature with Changes in MRT for Equal 31 Comfort 4.1.3 - Asymmetric Radiation and Comfort 32 4.2. Descriptions of Common Types of Radiant Systems 4.2.1 4.2.2 4.2.3 4.2.4 4.2.5 4.2.6 4.2.7 4.2.8 4.2.9 4.2.10 4.2.11 - Hydronic Floor Panels Electric Floor Panels Air Floors Hydronic Wall Panels Electric Wall Panels Hydronic Ceiling Panels (Metals or Plaster) Electric Ceiling Panels Miscellaneous Electric Systems Gas-Fired Radiant Ceramic Surface Infrared Units Gas-Fired Radiant Tube Infrared Units Electric Infrared Units 5.0 - CALCULATION OF DESIGN HEATING LOADS 5.1 - Standard ASHRAE Design Procedure 34 35 35 36 36 37 37 38 39 39 40 40 43 43 5.1.1 - Design Inside Air Temperature 5.1.2 - Room Temperature Gradients 5.1.3 - Wall, Ceiling, Floor Convection Coefficients 5.2 - Development of a Design Heat Loss Procedure for Radiant Systems 5.2.1 - Heat Balance on Room Surfaces 5.2.1.1 - q r - Radiant Exchange Rate 5.2.1.2 - q cv - Convective Heat Transfer 5.2.1.3 - qCQ< - Conductive Heat Transfer 5.2.2 - Heat Balance on Complete Room 5.2.3 - Comfort Equations 5.2.4 - Other Parameters Evaluated 5.3 - Comparison of Calculated Design Radiant Loads With the Standard 45 45 46 47 50 51 51 53 54 55 56 59 ASHRAE Design Load Calculation 5.4 - Test Case Calculation 60 5.5 - Radiant Panel Heating Systems Calculations 69 5.5.1 - Single Panel Radiant Heating Cases 5.5.2 - Effect Due to Infiltration for Radiant Panels 5.5.3 - Effect of Glass Distribution 5.5.4 - Changes in Wall, Floor, and Ceiling U-Factors 5.5.5 - Effect of Changes in Room Length and Width 5.5.6 - Changes in Room Height 5.5.7 - Changes in Outside Design Temperature 5.5.8 - Changes in Number of Panels 5.5.9 - Perimeter Panel System 5.6 - Comparison of Forced Air and Radiant Ceiling Panels 5.7 - Heated Floor Cases 5.8 - Infrared Heating Cases 5.9- U-Tube Infrared Cases 5.10 - Summary of Design Heating Calculations 6.0- DESIGN PROCEDURES 6.1 6.2 6.3 6.4 6.5 - Radiant Ceiling Panel Heating Systems Radiant Ceiling Panel Cooling Systems Heated Floor Systems High and Medium Temperature (Infrared Systems) Other Design Procedures 69 79 79 86 91 91 91 96 96 99 99 109 121 127 133 135 136 139 141 143 7.0 - SUMMARY OF MANUFACTURER SURVEY 145 8.0 - SYSTEM DYNAMICS 146 9.0 - RESEARCH NEEDS 147 9.1 - Convection Coefficients 9.2 - Air Temperature Stratification 147 147 '9.3 9.4 9.5 9.6 - Surface Emissivities Comfort During Radiant Temperature Asymmetry Radiant System Dynamics Heated Floor Systems 147 147 148 148 10.0 - REFERENCES 149 APPENDIX A - BIBLIOGRAPHY A-l APPENDIX B - ANNOTATED BIBLIOGRAPHY B-l A; B; c: D: E:1 F;1 G;1 H) i] J] K:) L; M;> N:I Load Analysis and Modeling Convection Coefficients General Comfort Conditions Thermal Comfort-Radiant Floor Panels Panel Heating and Cooling Infrared Heating Design Procedures Energy Consumption Transient Effects Instruments Controls Spot Heating and Cooling PENDIX C - LIS5TING OF COMPUTER PROGRAM C-l Program Listing Data Input File Listing APPENDIX D REPRODUCTION OF CHAPTER 8 FROM 1984 ASHRAE SYSTEMS HANDBOOK C-2 D-l SUMMARY The goal of this study was to obtain design data and relevant manufacturers data concerning the design procedures for radiant heating and cooling systems. A comprehensive literature search was conducted which resulted in an annotated bibliography with over 250 entries. This bibliography was subdivided into the following sections: load analysis and modeling, convection coefficients, comfort conditions, radiant thermal comfort, floor panels, panel heating and cooling, infrared heating, design procedures, energy consumption, transient effects, controls, and spot heating and cooling. The manufacturers survey resulted in identifying three commonly used categories of radiant heating/cooling surface temperature ranges. The low surface temperature range is 8O0F to 200oF for heating and 50oF to 70oF for cooling. The medium surface temperature range is from 700 to HOOoF and the high surface temperature range is from 1200oF to 2000oF. These surface temperature ranges identify the four commonly used systems for radiant heating and cooling: ceiling panel heating and cooling and floor heated panels operate in the low temperature range, U-tube infrared units operate in the medium temperature range, and modular gas-fired or electric infrared units operator in the high temperature range. Analysis of the above information indicated that the only reliable or appropriate design consideration would involve looking at the surface-to-air design process and not the means which is used to obtain the heated surface temperature. There are many variations or schemes used to obtain appropriate surface temperatures and it was not the object of this study to evaluate all of these schemes. Each manufacturer or designer has their unique method for obtaining a specific surface temperature. Descriptions and applications are provided for eleven of the most common configurations. These are: hydronic floor panels, electric floor panels, air floors, hydronic wall panels, elec- -1- heater surface temperature. and modular infrared) for typical ranges of many of the variables. electric infrared units. hydronic ceiling panels. ceiling height. -2- .trie wall panels. For each variation.5 ACH to -16% at 4 ACH. Design methods considering techniques for calculating loads. heater placement. gas-fired radiant tube infrared units. quantity of glass. and miscellaneous electric systems. sizing equipment and positioning equipment are presented for each of the common types of radiant heating systems. electric ceiling panels. The only variable which was found to have a significant effect on the difference between the actual design heat loss and the ASHRAE standard heat loss was the infiltration rate. number of heating surfaces. The actual design heat loss is less than the ASHRAE standard design heat loss. The design procedure for radiant cooling which is presented in the ASHRAE Systems Handbook was found to be adequate and is recommended for use. room size. heated floor panels. infiltration rate. outside air design temperature. convection coefficients. gas-fired radiant ceramic tube surface infrared units. gas-fired radiant ceramic tube surface infrared units. U-tube infrared. the required area of heater surface was calculated and the actual design heat loss for the radiant heating system was calculated and compared to the ASHRAE standard design procedure. A computerized technique was developed to relate heater surface temperature to the space heating requirements while maintaining the Fanger comfort constraints. The variables considered were: U-factors. The percent difference in these two design heating loads varied from -4% at 0. surface emissivities. Calculations were made for the four types of radiant systems (ceiling panel heating and cooling. and use of reflectors or deflectors on infrared units. a procedure for designing radiant heating and cooling systems has been developed. and positioning equipment. sizing equipment. A major effort of the project has been the preparation of an annotated bibliography of published sources of information for radiant heating/cooling systems. sizing equipment.OBJECTIVES AND SCOPE The goal of this project was to obtain a body of accurate and relevant data on methods of designing radiant heating and cooling systems. The data includes methods of calculating loads. From this material. additional research needed in order to improve the recommended methods for calculating loads. The study has focused on identifying all significant types of radiant heating and cooling systems by means of a literature search and analysis of appropriate available data and technical material. This procedure includes methods of calculating loads.0 . and system dynamics has been provided. positioning equipment. and positioning equipment. -3- .1. As a result of the preparation of this annotated bibliography. sizing equipment. producing comfort conditions. radiant heating systems increase the room surface temperature causing -4- . However. Convective types of systems using fans for delivering heated air which cause slight air pressure differences will tend to increase the air infiltration loss. that radiant heating systems produce higher floor. wall and glass temperatures due to the radiant heaters heating these surfaces and not the air. On the other hand. For this reason. A convective type of system such as shown in Figure 1 produces an environment where the air temperature is greater than the mean radiant temperature in the space.INTRODUCTION Convective and radiant heating and cooling systems have been used for many years in providing comfort systems in rooms occupied by people and/or materials. and thus producing greater heat losses to the surroundings. They claim that the room may be operated at a lower air temperature than if it is heated by a convective system because the radiant heat from the heater falls directly on the occupants. and the nature of the heat suppliers. on the other hand. floor and ceiling. The thermal environment within a room and its rate of heat loss are determined by the configuration and structural materials used in the walls. there is also the opposite factor. Proponents of radiant heating systems assert that these types of systems offer the potential for reduced heating unit sizes and reduced energy consumption. and thus there is no fundamental reason to expect them to be sized by the same technique or to require the same energy to produce identical levels of comfort. the amount of infiltration air forced through the room.2. produces an environment in which the mean radiant temperature (or "average" room surface temperature) is higher than the air temperature. A radiant heating system such as shown in Figure 2.0 . the infiltration air losses will be greater in convective than radiant heating systems. These two types of systems produce different comfort environments due to their nature of heat delivery or removal. FIGURE 1. FIGURE 2. CONVECTIVE TYPE OF HEATING SYSTEM 1 m J n < L Heating And/Or Cooling Panel EflH] I3LT15. RADIANT CEILING PANEL HEATING/COOLING SYSTEM -5- . In Chapter 25 of the ASHRAE Handbook of Fundamentals (1) a procedure is presented for determining the design heating load for a structure. In these systems. the surface temperature is -6- . There are two fundamentally different characteristics to be considered. One category is that of panel heating and cooling systems where the surface temperature can be called low and is in the range of 80 F to 200 F for heating and 50 F to 70 F for cooling. The fundamental objective of this project has been to determine if this ASHRAE Design Heating Load Procedure is applicable to radiant heating systems. full load hours. It is questionable whether some of the simpler procedures presented in Chapter 28 of the ASHRAE Handbook of Fundamentals (HBF) (1) such as degree day method. and these can be identified by the temperature range in which they operate. This project does not address the energy requirement calculation. System dynamics and thermal storage characteristics of the structure are important factors in answering this question.increased heat loss to the surroundings. Along with the total size of the heating system is the positioning of the individual heating units so that they provide uniform comfort conditions throughout the space. design calculations are made to indicate what is the expected maximum rate of total heat delivery which is to be expected from the heater. There are three general categories of radiant heating/cooling systems. The estimation of the energy used by a radiant heating system over a heating season is a separate and more complicated problem. First is the concept of sizing of radiant heating systems and second is estimating the energy required by radiant systems for providing comfort conditions over a heating season. Radiant heating systems also have the advantage of increasing the mean radiant temperature to which occupants are exposed and thereby allowing comfort at lower air temperatures.. or BIN method are applicable to radiant heating systems. For sizing. or in many cases. and in some cases a disadvantage. The second type of radiant system comprises the medium temperature range units which operate from about 700 F. Chapter 8 in Reference 2 presents a good description of these types of radiant heating or cooling systems as well as some design procedures for installing the systems.controlled in order to vary the quantity of heat being delivered or absorbed. They consist of gas or electric operated units placed at various locations throughout the space and are generally used for spot heating applications. The use of these types of radiant heaters is presented in Chapter 18 in Ref. These are illustrated in Figure 4. to 1100 F and consist of radiant tubes through which the products of combustion from a gas burner are circulated and then exhausted to the outside. and the application of the units is discussed in Chapter 18 in Reference 2. 2. Descriptions of these types of units are given in Chapter 30 in Reference 3. Descriptions of these units are given in Chapter 30 in Reference 3. Use of these units is depicted The gas fired units have the disadvantage of discharging the products of combustion inside the conditioned space. The advantage occurs when the units are properly sized and located. They have the advantage of exhausting the exhaust products to the outdoors rather than inside the structure. or electric current. walls or ceiling. Figure 2 illustrates a ceiling panel system and Figure 3 depicts a heated floor type of system. in Figure 5. and the temperature is maintained by circulating water. providing a higher mean -7- . air. for full area comfort heating. The third type of radiant unit is the modular high temperature infrared unit operating in the range of 1200 F to 2000 F surface temperature. One of the advantages. These units come in integral lengths which can be placed in specific patterns or in U-tube shaped units of different lengths. The controlled temperature surfaces may be in the floor. is the maintenance of comfort conditions when using radiant heating. I I 00 FIGURE 3 HEATED FLOOR TYPE OF RADIANT HEATING SYSTEM FOR BEDROOM AND BATH . Tube Heaters Medium Temperature U .Tube Heaters Medium Temperature feaJ FIGURE 4 MEDIUM TEMPERATURE RANGE RADIANT TUBE TYPE OF HEATERS -9- . CO u CD 4-1 Cfl CD Q O a CU U K) U < o IT) UJ a: cs -10- . the room air temperature gradient is negligible. It should be kept in mind. The disadvantage can occur if the radiant heat is concentrated to such a condition that the asymmetric temperature felt by the occupant is such that discomfort occurs in the space. These schematics were prepared from data such as that shown in Figure 7 which are some results from 1953 data at the ASHVE Laboratory in Cleveland. Typically. This occurs due to the fact that radiant systems provide higher surface temperatures than experienced in convective systems with very little air motion resulting in a more uniform air temperature distribution. For convection or forced air heating systems room. Ohio. Any design procedure that is specified must account for maintaining comfort and not creating severe asymmetric temperature conditions. insulation levels. no discomfort should be experienced by the occupants. For radiant heated and cooled rooms. and the air distribution system design and operation. a schematic is given of room air temperature gradients for forced air heating. heated ceiling panels. and heat floors. Convective heating systems will generally have air temperature gradients due to the higher temperature of the air brought into the space for heating purposes with a resultant higher air temperature at the ceiling than at the floor. Articles containing this type of data are listed under G-Panel Heating and Cooling in the ANNOTATED BIBLIOGRAPHY. that application of the available room air temperature gradient requires careful -11- . which then permits a lower air temperature for equal comfort conditions. In Figure 6. Another advantage claimed for radiant heating systems is the negligible air temperature gradient experienced by spaces using radiant sources rather than convective sources for heating. by satisfying the Fanger comfort equations [4] and limiting the asymmetric temperature to 9 F. air temperature gradients from 1/2 to 2°F per foot could be experienced depending on room size. however.radiant temperature for the occupants. SCHEMATIC OF AIR TEMPERATURE GRADIENTS FOR FORCED AIR HEATING. RADIANT CEILING HEATING. AND RADIANT FLOOR HEATING -12- .61° 68° 75° 82 61° 68° 75° 82 61° 68° 75° 82c FIGURE 6. ASHVE LABORATO RY DATA. °F FIGURE 7 MEASURED ROOM AIR TEMPERATURES FOR RADIANT HEATING AND COOLING SITUATIONS. \ AUST = 85 F \ cu fa UJ 4-1 I . JULY 1953 .T—r T I Ceiling Ceiling Panel Cooling.C 60 •H ID a Heated Floor Floor a t 85FJ AUST = 65 F Floor 65 70 75 80 85 ROOM AIR TEMPERATURE.Ceiling at 65F. This exfiltrated air will be at a higher temperature in the convection type systems thereby creating a larger heat loss than experienced in radiant types of systems. and manufacturing or industrial situations . and warehouses. What has been investigated in this project is a system design procedure for radiant heating and cooling. Radiant heating systems are used in many types of applications such as offices. manufacturing. the selection of the type of radiant system to be used (ceiling panels. and industrial situations. With a temperature gradient in the room. The medium and high temperature infrared systems are generally found in warehouses. it is expected that convection types of heating systems will have larger design infiltration losses than radiant systems. The evaluation of the energy requirements for radiant heating systems has not been considered. Because of this. offices. warehouses. hospitals. and -14- . This is another expected benefit of radiant heating types of systems. These general types of applications are not meant to be restrictive since each application should be addressed individually by weighing the advantages and disadvantages of each type of system. very often radiant floor panels are used.consideration of all of the parameters under which the data was measured and collected. All data cannot arbitrarily be applied to any situation. infiltration air will enter at the bottom of the space and exfiltration will occur at the top of the space. For hospitals and offices ceiling panel radiant systems are typically used. the infiltration heat loss is greater than when a gradient does not exist. floor panels. During the heating mode. The system design procedure involves the estimation of the design heating or cooling load. homes.2. Additional details on applicability are given later on in this report in Section 4. U-tube modular units or infrared modular units) which is partially based on the allowable heater surface temperatures for the application considered. For homes. : -15- . &-. a literature survey has been conducted and is included as Appendix A and an Annotated Bibliography is presented as Appendix B.the positioning of the heaters in the space. In addition. Some of the terminology concepts are important because names have been commonly associated with various phenomena or items which are restrictive when they need not be.T 4 ) (1) where: a = Stefan-Boltzman Constant.1 .Radiation.1713x10"8 Btuh/ft2 R 4 A = area of one surface F = geometrical factor relating shape and orientation of the surfaces ^1-2 ^ net exchange of radiant heat between the two surfaces 3.Convection.1.3. The motion of the fluid may be entirely the result of differences of density resulting from the -16- . Although rate of energy emission is independent of the surroundings. Convection involves the transfer of heat by mixing one portion of fluid with another.1 .Radiation. its microscopic arrangement and its absolute temperature. net energy transfer rate depends on the temperatures and spatial relationships of the surface and its surroundings and can be expressed for two black surfaces as: qi_2 = orAF (T4 . 3.0 . Amount and characteristics of radiant energy emitted by a quantity of material depends on the nature of the material.DEFINITIONS AND TERMINOLOGY This section of the report presents some definitions and terminology that are used throughout the remainder of the report. 0.2 . Convection and Conduction 3. The radiation energy transfer process is the consequence of energy carrying electromagnetic waves emitted by atoms and molecules resulting from changes in their energy content.1. T 2 ) (2) where: q = exchange of convective heat between two surfaces h = convection heat transfer coefficient A = surface area 3.3 .2 . Conduction in a homogeneous opaque solid is the transfer of heat from one part to another under the influence of a temperature gradient without appreciable displacement of the particles. or the motion may be produced by mechanical means such in forced convection. K A (TX .temperature difference. For steady state one-dimensional conduction heat transfer. Conduction involves the transfer of kinetic energy from one molecule to an adjacent molecule.T 2 ) X (3) R where: q = exchange of heat by conduction from one surface to another K = thermal conductivity of the material X = thickness of the material A = area perpendicular to the flow of heat R = thermal resistance of the material C = thermal conductance of the material 3.Conduction.T 2 ) T]_ .1.[5] The general equation for convection heat transfer is q = h A (T-L .Infrared Ranses The thermal radiation emitted by a surface encompasses a range -17- . as in natural convection.T 2 q = C (T]_ . the following equation applies. c o Spectral distribution 2c o o ~ w> ro vt Wavelength Figure 8 Spectral Distribution of Thermal Radiation [6]. Also delineated there are the various wavelength regions of thermal radiation. -18- . nonuniform distribution of monochromatic (single-wavelength) components. Directional distribution Figure 9 Directional Distribution of Emitted Radiation [6]. Emitted radiation consists of a continuous. An example of this is shown in Figure 8 where the magnitude of the radiation varies with wavelength. The relationship between temperature and wavelength for the peak radiated energy is illustrated in Figure 10. The magnitude of the radiation at any wavelength and the spectral distribution vary with the nature and temperature of the emitting surface. It is also important to understand that a surface may emit preferentially in certain directions creating a directional distribution of emitted radiation as illustrated in Figure 9 [6].of wavelengths. Wavelength Relationship Radiant heaters used in comfort heating applications operate in what is generally classified as the far-and middle-infrared region: 2 to 8 microns wavelength and 85 F to 2000 F surface temperature. Medium and High Temperature Radiant Sources Commonly available radiant heating systems for HVAC applications exist in three general temperature ranges.WAVE LENGTH (IN MICRONS) 30 20 IS FAR INFRA- 10 -O'F RED 8 6 SOO'F 5 4 MIDDLE 3 INFRA- -|0O0oF RED a • -3OO0°F i. The low ranpe is from 80 F to 200 F and consists of hydronically or electrically -19- .3 Low.8 NEAR INFRARED 6000°F VISIBLE LIGHT ULTRA VIOLET Figure 10 Temperature . 3.s i . + T 4 F p . 3. -. T2.heated panels or surfaces usually placed in the ceiling or walls or as part of the floor: The medium range units operate between 700 F and 1100 F and usually consist of vented gas-fired tubes placed near the ceiling or roof in a structure. The hiph temperature radiant units operate in the range of 1200 F to 2000 F and usually are nonvented gas-fired units or electrically heated units.Radiant Temperature Asymmetry The difference between the plane radiant temperature of the two opposite sides of a small plane element. MRT 4 = T^Fp. The MRT can be determined from the following equation.Mean Radiant Temperature (MRT) The temperature of a theoretically conceived isothermal black enclosure in which an occupant would exchange the same amount of heat by radiation as in an actual nonuniform surface temperature environment. For -20- .6 feet above the floor at the person's position..T n = surface temperatures surrounding the occupant in a room *p-l» *p-2» "" F p-n = geometrical factor relating shape and orientation between a person and the surrounding surfaces. For vertical heating or cooling panels it refers to a small vertical element 3. This is a new term used to describe the asymmetry of a radiant environment. n (4) where: T^.5 .4 .! + T 4 F p _ 2 + --. 3. -. 3. F e _2.+ F e .6 .T n = temperatures of surfaces surrounding the element Fe_^. It is given by the following equation.F e _ n = geometrical factors from the plane element to the specified surface.Operative Temperature (tQ) The uniform temperature of a theoretically conceived enclosure in which an occupant would exchange the same amount of heat by radiation and convection as in the actual nonuniform surface temperature environment. The plane radiant temperature (T_r) can be calculated [7] from T pr = F e-lT 4 + F e . 2 T 4 + --. (hc x t a ) + (hr x MRT) tG = = (7) hc + hr where: h c <•= the convective heat transfer coefficient for the occupant [1] h r = the radiant heat transfer coefficient for the occupant [1] t a = ambient air temperature -21- . n T 4 (5) where: T^. T2. Radiant temperature asymmetry is then defined as T prl " Tpr2 <6> where the subscript 1 refers to one side of the plane element and the subscript 2 refers to the opposite side.horizontal heating or cooling surfaces it refers to a small horizontal element at the same position. -. + A N T N AUST (9) A l + A 2 + •• + A N where: A l> A2> "" ^n = areas °f surfaces not supplying external heat to a room T^.7 .Average Unheated Surface Temperature (AUST) The area-weighted temperature of the surfaces in a room which are not acting as suppliers of external heat to the room.h r (MRT-ta) .T n «= temperature of surfaces not supplying external heat to a room 3. T 2 .Effective Radiant Flux (ERF) This is the net radiant >heat exchanged per unit area at the ambient temperature t a between an occupant represented by a hypothetical surface and all enclosing surfaces and directional heat sources and sinks. The frequently used parameters are: t a = ambient air temperature RH = relative humidity -22- . It is given by AlTi + A 2 T 2 + -.3.Comfort Conditions Several parameters are used to identify when a human occupant is exposed to what are commonly called comfort conditions. ERF .8 .9 .h c (tQ . -.t a ) (8) 3. It is given by the following equation [1]. ERF is the net radiant energy per unit area received by the occupant from all surfaces and sources whose temperatures differ from the ambient air ta. 10.t oa ) where: V-i '«=» design U value for each component of the room (walls. floor) AJL = area of each of these individual components -23- (10) . In addition. These are defined as follows.1 CFMI (75 . MET = the metabolic rate for the occupant which is a function of their activity level CLO = the thermal resistance of the clothing being worn by the occupant The above variables have been combined into a set of equations • [4] which can be solved yielding values for the comfort range. glass. ceiling.10 .Design Heat Loss Values There are four design heat loss values which will be calculated during this analysis. 3. Three of these are different from what is taken as the "standard ASHRAE design heat loss".3.V — relative air velocity MRT = mean radiant temperature There are two other items which affect the values of these parameters when comfort is concerned. These are discussed later in Section 5. sets of charts have been developed [1] which are applicable for the most common range of conditions.t oa ) +1.ASHRAE Standard Heat Loss (HLD) HLD - SiUjAi (75 .2. 3.1 . 2 .t oa ) (13) where: ta„ = The room air temperature at the ceiling level where exfiltration occurs. It is evaluated using a specified air temperature gradient based on a reference height of 5 ft.S U t k± (ta .t oa ) + 1 . 1 CFMI (t ag .t o a «= outside design air temperature CFMI = estimated design value of infiltration air 3.10.t oa ) + 1.Conduction Design Heat Loss with Room Air Temperature Gradient (HLCG) HLCG .Conduction Design Heat Loss (HLC) HLC ~ S ± C t A £ (t si .10.t oa ) +1.t oa ) (11) where: t a = design room air temperature based on comfort conditions being met at the center of the room 3. glass.10.3 .1 CFMI <ta .1 CFMI <ta .t oa ) (12) where: C^ = The conductance for the room component (wall.Actual Design Heat Loss (HLA) HLA .S ± Ci Ai (t si . floor) from the inside surface to the outside air. t si = It excludes the inside convection coefficient.4 . *-he surface temperature of the room component 3. ceiling. -24- . t a .1 .BACKGROUND 4. t m r t .Environment variables Fanger's Comfort Equation is represented by the following two equations. Pa.1. 4. The Fanger Comfort Equations are based on a logically derived heat balance equation for the occupants thermal equilibrium and on observations that during a state of comfort defined by neutral temperature sensation.Relationship Between Radiant Heating/Cooling and Comfort Conditions When studying the use and performance of radiant heating and cooling systems it is important to understand the inter-relation between the various parameters of comfort and their influence on the sizing and location of radiant units. a unique relation exists between level of activity.Fanger's Comfort Equations Fanger's study generalizes the physiological basis of comfort and allows comfort for most activity levels to be predicted analytically in terms of the environmental parameters presented in 3. The comfort equation contains the following grouped variables: Ic]_t fcl " A function of the type of clothing M —-. and evaporative loss from the body [1].DEFINITIONS AND TERMINOLOGY. The one which is most suited for this investigation and which has been widely accepted as providing meaningful results is the work by Fanger [4]. Many comfort studies have been done over the years and many of these are presented in the ANNOTATED BIBLIOGRAPHY in Appendix B.0 . skin temperature.-IJ'.0 . v .1 .4.A function of the type of activity A Du v. -25- . 0 .35 [43.Pa) A Du Du M 0.M M (l-n) A .0014 A - Du M 0.(tnrt + 273)4] + f cl h c (fccl .ta)] A Du Du (15> where: M = metabolic rate A Du = DuBois body surface area rj = external mechanical efficiency P a = water vapor pressure of ambient air t a = ambient air temperature fc1 = the ratio of the surface of the clothed body to the surface area of the nude body t c ^ = clothing surface temperature tj^-j.P a ) -0.0 .50] Du M .0.0 .32 A Icl [ A Du (l-q) M (1-r.0.50.ta) M tcl = 35.) .) .0014 (34.= mean radiant temperature -26- .61 (14) (34.0 ] .0 .7 M (1-IJ) .18 .Pa] A Du Du M M 0.0.42 [ (l-»7) .35 [43.0.0.0 .8 f cl [t cl + 273) 4 .0.0023 (44.0023 A (44 .0.t a ) A Du 3.61 (l-i.42 [ A A Du M (1-r.) .4 x 10.0.P J . L = thermal resistance of clothing The ASHRAE [1] comfort envelope is shown in Figure 11 but applies only for sedentary (1 Met =58. -27- . "C 30 Figure 11. Acceptable Ranges of Operative Temperature and Humidity for Persons Clothed in Typical Summer and Winter Clothing at Light to Sedentary Activity [1]. Figures 12. 13 and 14 depict comfort lines or curves / 20 / / • / 15 7 | i^ % S 1 /' f? - : ~ 0 -5 - -10 20 25 OPERATIVE TEMPERATURE.2 W/m2) and slightly active (<1. activity and environmental parameters given by Fanger's Comfort Equation and shown in these figures or comfort charts has been confirmed by studies of the individual parameters [8-13]. generally a function of air velocity IC. through various combinations of variables in order to create comfort for constant values of some of the other variables. The quantitative influence of clothing. normally clothed persons at low relative air velocities when the MRT equals the air temperature.2 MET).h c = convective heat transfer coefficient. "C MEDIUM ACTIVITY 2 mel MEDIUM ACTIVITY 2 mel LIGHT CLOTHING I c | = 0.5clo MEDIUM CLOTHING Iei=I.. *C • - 30- " - 20 1—I—|—I—p 25 AIR TEMPERATURE = MEAN RADIANT TEMPERATURE. I 3° \f ^S* ^ Ysk s> Si g is- CO fe Jta 10- - TO lu» J^T° • 5- . "C COMBINED INFLUENCE OF HUMIDITY AND AMBIENT TEMPERATURE [ 1 ] -28- . "C HIGH ACTIVITY 3 met HIGH ACTIVITY 3 met MEDIUM CLOTHING Ic|S|0ClO LIGHT CLOTHING IcI=0.5clo r&S 25- JS ••tir S^ O S / jr$^\y^ S rA'S' ' uf Ul a.MEAN RADIANT TEMPERATURE. ' C AIR TEMPERATURE = MEAN RADIANT TEMPERATURE.SEDENTAHY I mel AIR TEMPERATURE = MEAN RADIANT TEMPERATURE. *C FIGURE 12 -I—PI—|—P 1 21 AIR TEMPERATURE .Oclo 10 AIR TEMPERATURE = MEAN RADIANT TEMPERATURE. Âo y* -g^^ ..^ o2i^--^' 13_ • * " * ^ X>^ J^>^ i -T 10 ^^ [ i 1 1 1 1 1 1 1 1 1 1 1 1 1 I IS 20 25 30 AIR TEMPERATURE = MEAN RADIANT TEMPERATURE. ...V ^ / \ / ^ vl ^ SEDENTARY 1 met / / / • o AVI* IING » c l ' 0..AV ..... 20 25 30 AIR TEMPERATURE. 35 • • .. MEDIUM CLOTHING \ Id'IOdo • 1 i!\ . ..C I I I I—I I I I IS 20 25 30 35 AIR TEMPERATURE. I . . IS .. I I IO 10 FIGURE 13 . *C 30 35 / N / \ ^ s \ / < V& V / / / * / / MEDIUM ACTIVITY 2 mel .5 d o / S / • 1 ft/ V • • • \ .... 5 . 10 . • IS 20 25 AIR TEMPERATURE. .. 10 V 15 20 25 AIR TEMPERATURE. 3 0 ... "C 35 . *C COMBINED INFLUENCE OF MEAN RADIANT TEMPERATURE AND AIR TEMPERATURE [ 1 ] -29- . .. . LIGHT CLOTHING IcfOSeto IC|*I. the comfort lines are curves through different combinations of mean radiant temperature and air temperature that provide thermal comfort.41.2- *u *"/ §0B- > | 7 06- ? 1 \ 7 / J / 1 1 1 1 1 1 1I (J i i i i */ of f '§^ \ /J i / 1 \1 >. The six charts apply to six different combinations of activity and clothing. In Figure 12 the comfort lines are curves through different combinations of ambient temperature and humidity that provide thermal comfort. i / J i i i i i 10 13 20 23 AIR TEMPERATURE o MEAN RADIANT TEMPERATURE. IO- hi/ 7 y / / . In Figure 13.i i/ i. Combined Influence of Air Velocity and Ambient Temperature [1].I I i i i 1. *C 90 Figure 14. where the air temperature equals mean radiant temperature.'C I I 1 I | I I I I | I IO 13 20 23 AIR TEMPERATURE " MEAHRADIANTTEMPERATURE. -30- The six charts apply to six . different combinations of activity and clothing at 50 percent relative humidity. activity level and clothing level. For the other situation with medium activity and light clothing the temperatures are equal at 69. 4. if the MRT = 77 F the room air temperature needs to be at 63.0 clo with the relative humidity at 50 percent.5 F which is a common situation in convective heating systems. 12-14.6 F. In this table.8 F. For practical application of these charts. These are illustrated in Table 1. Notice that the air temperature should be lowered to 72 F for comfort if the MRT = 80.Changes in Air Temperature with Changes in MRT for Equal Comfort Some comfort condition examples for radiant heating applications have been worked out in order to illustrate how the ambient air temperature should be changed for different mean radiant temperatures at various activity and clothing levels. In Figure 14. MRT. The two charts apply for persons wearing 0. This illustrates the large changes in room air temperature which are needed in order to provide comfort for changes in MRT. the marked values are when the mean radiant temperature is equal to the air temperature. taking room use into account.2 . first estimate the activity level and clothing quantity.5 and 1. For sedentary activity and medium clothing these temperatures are 76. In this situation. -31- .1.6 F for sedentary activity and light clothing to 61. This value ranges from 80. the comfort lines corresponding to five different activity levels are curves through different combinations of relative air velocity and ambient temperature which provide optimal thermal comfort. RH and velocity) that provide thermal comfort can then be found from Figs.2 F for medium activity and clothing.5 F for comfort. The combination of the four environmental parameters (ta. The strong effect of activity and clothing level can be illustrated by looking at the temperatures when they are equal. F F • • 80.8 71. then the lowest acceptable radiant -32- .3 75.8 81.0 56.5 F MRT.6 72 78.8 66.9 73.8 82. F ta.2 73.8 71.2 80.6 59.4 78.6 72.6 • 61.5 75.8 69.0 75. 75 50 67. F ta.6 67.5 51.6 62.6 64. Asymmetric Radiation and Comfort.5 68.3.0 59. Radiant temperature asymmetry was defined in 3.2 62.8 86.7 77 63.4 66.6 53. F MRT. A figure presented by Fanger [14] is shown as Figure 15 delineating radiant temperature asymmetry for heated ceilings and walls and for cooled walls and ceilings.6 65.3 80.5 75.7 63.4 77 74.7 69.2 71.4 64.1 62.6 74.4 79.6 85.9 66.4 63.2 50.4 72.6 77 71. If 10 percent dissatisfaction is acceptable.TABLE 1 Comparison of Room Air Temperature and MRT for Comfort Conditions at Different Activity and Clothing Levels with 30% Relative Humidity and 0.4 4.8 61.6 96.4 55.1 69.2 m/s Relative Velocity Sedentary Activity Sedentary Activity Medium Activity Medium Activity Light Clothing Medium Clothing Light Clothing Medium Clothing MRT.2 69.7 86 68.1.1 84.8 62.2 71.0 62.2 79.5 76.4 57.2 57.9 70.4 60.2 64.0 ta.9 • • • 93.8 68. F ta F MRT.3 53.8 78.4 75.5 66.2 89.6 71.8 60.3 64.6 80.6 81.2 83.0-DEFINITIONS AND TERMINOLOGY and recent work has suggested some values for acceptable radiation asymmetry for comfort. 68 87.6 68 79 68 • • 82.6 76. 100 60 40 •a ai •H >H 20 to •rl 4-1 CO as 10 - CO •H « n (U o <u 5 4 3 Warm Wall 5 10 15 20 25 30 35 Radiant Temperature Asymmetry.temperature asymmetry is approximately 12 to 13 F. °C Figure 15 Percentage of People Expressing Discomfort Due to Asymmetric Radiation [14] Additional work done by Olesen and Nielsen [7] resulted in the values shown in Figure 16 for vertical radiation cooling. The ISO standard [15] suggests 9 F. (15K) -33- This allows about a 25 F . the following descriptions were prepared.Description of Common Types of Radiant Systems A survey of manufacturers and designers was conducted in order to identify the commonly used types of radiant systems. K Figure 16 Percent Dissatisfied in Spot Cooling Due to Vertical Radiation [7] radiation asymmetry for 10 percent of the occupants being dissatisfied. Generally. The true value of this quantity in actual applications is difficult to establish since it depends on the type of radiant unit. the amount of reradiation and the angle factor between the occupant and the radiant unit. Additional research is required in this area.« - 60 40 CD 20 nf - - • : CO CO 10 cu bO ca 6 m 4 - • 2 - m •U c CU o 1 cu P4 0 2 4 6 8 10 12 14 16 18 20 22 24 Radiant Temperature Asymmetry.2 . an applications matrix was developed from the manufacturers information and the previously given descriptions of the various types of radiant heating and cooling systems. 4. From the results of this survey In Table 2. its location relative to the occupant. about a 9 to 15 F radiant temperature asymmetry appears to be acceptable. it has not been calculated in this study. the location of furnishings in the room. For these reasons. Ten -34- . It should be kept in mind however that there are other applications and characteristics which are not identified here. For comfort applications. 4. although metal pipes or copper tubing can also be used. 4. plastic pipe is used. the floor surface temperature is limited to a maximum value of 85°F.2.2 . and light commercial structures. the floor surface temperature is generally limited to a maximum value of 85°F. Typically.2.Electric Floor Panels This type of radiant heating system (see Figure 3) consists of electrical heating elements imbedded in concrete floors. factories.Hydronic Floor Panels This type of radiant heating system (see Figure 3) consists of pipes imbedded in concrete floors with heated water being circulated through the pipes.1 . They are commonly applied in residences. It is not possible to list all conceivable applications and operating traits for all radiant systems. Hydronic floor panel systems are well suited for applications where a large change in load does not occur in a short time span. Electrically heated floor systems provide a uniform source of heating and require no auxiliary systems unless forced ventilation is required for the space. They can be -35- . For comfort applications. garages. Their transient responses are slow due to the high thermal mass of the concrete floor. Hydronic floor systems provide a uniform source of heating and require no mechanical air circulation unless forced ventilation is required. They can be zoned for various types of load situations and create a minimal amount of noise during operation. although new lower mass floors have helped to alleviate this problem.types of radiant systems are compared and some of their general characteristics and typical applications are indicated. warehouses. the floor surface temperature is limited to a maximum value of 85°F.Hydronic Wall Panels Hydronic wall panels can be modular metal panels with tubing connected to the backside. 4. providing ventilation if required.zoned for various types of load situations and create no noise during operation.3 . factories. garages. They are commonly applied in residences. warehouses. the heated air from a furnace is circulated through passageways in the floor (wood construction or concrete with imbedded tile) .2.Air floors In air floor types of radiant heating systems. although new lower mass floors have helped to alleviate this problem. For comfort applications where people are present. warehouses. and light commercial structures.2. They are easily controlled with shielded thermostats. This source of air can also be used for These systems can be zoned for various types of load situations and normally create only a minimal amount of noise during operation. Air heated floor systems are well suited for applications where a large change in load does not occur during a short time span. Electric floor panel systems are well suited for applications where a large change in load does not occur in a short time period. They are normally controlled with shielded thermostats. and portions can be sequenced for more efficient energy usage. They are commonly applied in residences.4 . 4. Their transient response is slow due to the high thermal mass of the concrete floor. Air heated floor systems provide a uniform source of heating and do require mechanical circulation of heated air. and light commercial structures. or tubing connected to the wall surface and covered with -36- . office buildings. where tubes on the backside of the ceiling panels. The surface temperature would have to be limited if it can be contacted by people. Hydronic wall panels are used in place of ceiling panels when the panel location in the ceiling would interfere with lighting fixtures or some type of required suspensions from the ceiling.plaster. Generally more heated panel area will be required than with ceiling panels due the wall location. industrial plants and sports facilities.Hydronic Ceiling Panels (Metal or Plaster1) Hydronic ceiling panels (see Figure 2) can be made up of modular metal panels laid in the ceiling. These units have the same characteristics and features as the hydronic ceiling panels. 4. Electric wall panels would be used in lieu of ceiling panels when the panel location in the ceiling would interfere with lighting fixtures or required suspensions from the ceiling. industrial plants and sports facilities. These units have the same characteristics and features as the electric ceiling panels. Electric cable can also be attached to the wall and covered with plaster to accomplish the same type of heating.5 . carry circulated water for heating or cooling or tubing attached to the ceil-37- . 4. These units are applied in hospitals. The surface temperature would also have to be limited if it can be contacted by people. office building.2.Electric Wall Panels Electric wall panels are available in various sizes for different types of applications and are composed of heaters located between a wall surface material and insulation on the back of the panel. Generally more heated panel area would be required than with ceiling panels. These units are applied in hospitals.6 . or if the ceiling is too high for practical application.2. office buildings. the entering water temperature can be no lower than the room air dew point. These types of systems are generally applied in hospitals. The metal panel systems will respond quickly to load changes in the space. For heating. the panel surface temperature is generally in the range of 120°F to 180°F depending on ceiling height. These are used for heating only applications and generally operate with surface temperatures between 120°F and 180°F. industrial plants. When cooling is done. These units can be zoned for various types of load situations arid can be used in conjunction with other types of heating/cooling systems. Electric ceiling panels provide a uniform source of heating or cooling and radiate to the floor and walls maintaining them at comfortable temperatures. They can also be used in conjunction with other types of heating systems within a building. and sports facilities. schools.Electric Ceiling Panels Electric ceiling panels (see Figure 2) are available in various sizes for different applications and are made up of various types of heaters located between a ceiling surface material and insulation on the back of the panel. and the latent load in the space must be removed with a separate system.2. 4.7 . air terminals. Electric radiant celing panels are applied in hospitals. Hydronic ceiling panels provide a uniform source of heating or cooling and do radiate to the floor and walls to maintain them at a comfortable temperature. These systems can provide both heating and cooling and can use any source of energy since water is the circulated heat transfer fluid.ing substructure covered with architectural plaster. Their transient response is generally good due to their low thermal mass. These units can be zoned and controlled sequentially in order to accommodate various load situations. office -38- . 4. aircraft hangars. warehouses. schools. arenas. Units are available with different types of reflectors or lens. 4. They are easily installed and do not require construction changes within the building. They are not suited for applications. These devices are generally used for retrofit types of applications or where some additional heating is necessary. which concentrate and direct the radiant energy into suitable patterns. One form is an electric carpet that is composed of electrical heating wires. a metallic grid is placed at the surface to enhance performance of the unit.9 . industrial plants and sports facilities. They are most commonly used in occupancies where a large room volume is present with high ceilings.Miscellaneous Electric Systems There are several forms of unusual types of radiant heating systems available. air terminals. and combustion occurs at the surface of the refractory burner.8 .buildings. and auditoriums. In some instances. There are also cloth wall hangings or drapes of similar construction involving electrical heating wires.2.Gas Fired Radiant Porous Refractory Surface Infrared Units These radiant heating units (see Figure 5) supply a mixture of air and gas through a porous refractory material. Units used for total building heating typically operate with a surface temperature in the range of 1500 F (815°C) to 2000 F (1094°C) and are self-contained and use shielded air-source thermostats for control. where combustible or explosive fumes are present. assembly areas. and where the additional -39- . Common applications include factories.2. They are well suited where people or materials are to be heated and it is not necessary to heat the surrounding air. These units are unvented and place the products of combustion (mainly carbon dioxide and water vapor) into the space being heated. designs. Some units use reflectors to concentrate and direct the radiant energy to areas where it is needed. 4.2. gas and air are burned in a combustion chamber and the products of combustion are forced through a tube providing a radiant energy source. the tubes generally vary in temperature from 900°F to 500°^ al°ng the length of the tube between the combustion chamber and the exhaust vent.10 . Various types of reflectors located above the tube are used to concentrate and direct the radiant energy to the lower levels of the structure where it is needed. warehouses. quartz tubes. 4. The major advantage of these units over other types of gasfired infrared units is their larger radiating surface and the fact that the combustion products are vented from the space. and other total heating applications in high volume spaces.11 .2. aircraft hangars. (see Figure 4 ) . The products of combustion are then exhausted to the outside and not into the space being heated. arenas. They do not require combustion air for operation and therefore do not have to dispose of combustion products. They are available as U-shaped or linear tubes for versatility in The radiant heating tubes are generally about 4 inches in diameter. lamps or panels for delivering infrared heating to large volume spaces.Electric Infrared Units Electric infrared units (see Figure 5) use metal rods. auditoriums. Panel type units have a surface temperature between 200°F and 1100°F while the metal and quartz tube units -40- . garages.Gas-Fired Radiant Tube Infrared Units In this type of gas-fired infrared system. These units provide a uniform source of radiant energy at a low intensity level and are generally used in factories.moisture due to the combustion products will be detrimental. and during operation. assembly areas. -41- . Quartz lamps generally operate at about 4000°F surface temperature.operate at a surface temperature between 1500°F and 1800°F. arenas. These units are designed to be used for spot heating applications and large volume spaces where people. warehouses. or surfaces are to be heated rather than air. swimming pools. They are normally controlled from shielded thermostats in the space. and auditoriums. aircraft hangars. objects. These include factories. m E 2 /miavncNS WORK Type of Radiant System Integral Exhaust Total Cooling With Response Surface Capacity Venting or Spot System Tenp. Construction Tire or Required Keating F Add-On Possibility . Condensation to be Considered Applications Residential Industrial Warehouse Garage Comrerclal Sports School Hospital Office Facility Hydremic Floor 85 Integral slow total No No No X Electric Floor 85 Integral slow total No No No X A1r Floor 85 integral medium total No No No X Hydremic Wall 100 iiitegral msdlum total Yes No If Cooling X X Electric Wall 100 Integral (redium total No- No No X X Hydremic Celling 55-230 Add-on good total Yes No If Cooling X X X X X Electric Celling 120-200 Add-on good total or spot No No No X X X X X Ceramic Infrared 1500-1700 Add-on good total or spot No Yes I f not vented X X X X Tube Infrared 700-1200 Add-on good total or spot No Yes I f not vented X X X X Electric Infrared 1100-4000 Add-on good total or spot No No X X X X No X X X X X X X '• X X . Standard ASHRAE Design Heat Loss Procedure In Chapter 25 of Reference 1. prepare the following information about building design and weather data at design conditions. Winter climatic data can be found in Chapter 24. but has not been validated with experimental data. the energy required to warm the outdoor air to the space temperature must be provided by the unit. wind speed and the temperature difference between indoor and outdoor air. as well as any cold walls. The procedure developed for this project is based on the best available information for radiant and convective exchange. 3. The principle for calculation of this load component is identical to that for infiltration. 8. From that source. GENERAL PROCEDURE To calculate a design heating load.0 . some reduction in infiltration may occur.1 . 5. Select unit values and compute the energy associated with infiltration of cold air around outside doors. Compute heat transmission losses for each kind of wall. for inside walls. Estimate temperatures in adjacent unheated spaces. These unit values depend on the kind or width of crack. using inside dimensions. followed by a procedure developed specifically for this project.) 9. what is commonly referred to as the ASHRAE standard procedure is presented. 5.CALCULATION OF DESIGN HEATING LOAD This investigation is directed at providing a design procedure for radiant heating and cooling systems. glass. 1. floors or ceilings next to unheated spaces. the following general procedure has been reproduced. if these are next to unheated spaces. 4. 6.5. nonbasement floors and ceilings. Select the indoor air temperature to be maintained in each space during coldest weather. Select outdoor design weather conditions: temperature. This procedure will be presented first. ceiling and roof in the building by multiplying the heat transfer coefficient in each case by the area of the surface and the temperature difference between indoor and outdoor air or adjacent unheated space. 2. (See Chapter 22. If no mechanical exhaust is used and the outdoor air supply equals or exceeds the amount of natural infiltration that can occur without ventilation. glass and roof next to heated spaces. An alternative method is to use air changes. -43- . and the roof if it is next to heated spaces. wind direction and wind speed. The major concern is whether or not the ASHRAE standard heating load design procedure can be used for radiant systems. and a discussion on the differences between these techniques and some other techniques found in the literature. the unit must also provide for natural infiltration losses. Select or compute heat transfer coefficients for outside walls and glass. If mechanical exhaust from the space is provided in an amount equal to the outdoor air drawn in by the unit. These determinations can be made from building plans or from the actual building. windows and other openings. 7. floor. Compute heat losses from basement or grade-level slab floors using the methods in this chapter.' When positive ventilation using outdoor air is provided by an air-heating or air-conditioning unit. Determine net area of outside wall. In buildings with a sizeable and reasonably steady internal heat release from sources other than the heating system. =2808K. 11. LW -44- . ceilings. 4 to assist in determining TD Floors Above grade On grade Below grade . for attic temperatures.Summary of ASHRAE Standard Design Heat Loss Calculations rX) Healing Load Roofs. Description Equation -•-Chapter 23.For crawl space temperatures. -See Table 5 TD Perimeter of Slab q = {/• A • TD -Use Fig. ssee Eq. See Chapter 22 for estimating methods for inIfiltration qs = 1200K* Af -Humidity ratio difference q. (1). a table which is given in [1] that summarizes the typical load calculations is reproduced below. (2) -Area calculated from plans -See Table 3 q = U' A • TD -*. Consider using pick-up loads that may be required in intermittently heated buildings or in buildings using night thermostat setback. The sum of the transmission losses or heat transmitted through the confining walk. In addition. air and material contents to the specified temperature. see Eq.A.TD ~[~ ^ ITemperature difference between inside and outside design dry bulbs. Table. represents the total heating load. Chapter 24. glass and other surfaces.. Tables 3 and 4 q = U . glass Walls below grade Reference. plus the energy associated with cold air entering by infiltration or required to replace mechanical exhaust.Use Fig. For tem*1peratures in unhealed spaces. 12. 4 to assist in determining TD -•-See Table 4 Infiltration and ventilation air Sensible Latent J—- _JVolume of outdoor air entering building. see Eq.10. (4) q= U-A-TD q = F2 • P . TABLE 3 . compute and deduct this heat release under design conditions from the total heat losses computed above. ceiling. floor. Pick-up loads frequently require an increase in heating equipment capacity to bring the temperature of structure. walls. the value used for the inside design temperature in this analysis has been set at 75 F. It is common practice to select the inside design dry bulb temperature at 75 F in most localities in the United States. For higher ceiling/roof rooms or spaces. 5.The components of the ASHRAE standard design heat loss (HLD) are transmission losses and infiltration losses.1 CFMI (75-toa) (17) or ACHxV 1. 5.t oa ) (18) 60 where ACH is the number of infiltration air changes per hour and V is the volume of the space.1 x ( ) x (75 . but is usually not incorporated into the design heat loss calculation.Design Inside Air Temperature. this gradient can affect the design heat -45- . The transmission losses are: S UjAi (75 . this is done without accounting for the comfort constraints previously described.t oa ) (16) and the infiltration loss is 1.1.0 -DEFINITIONS AND TERMINOLOGY. a small temperature gradient in room air may exist as discussed. Generally. As indicated in the section 3.1. This temperature is used in both the transmission loss and infiltration loss calculations and a choice of this value will affect the design loads for the space in proportion to the temperature difference between the inside and outside at design conditions.Room Air Temperature Gradients. For rooms which are eight to ten feet high.1 .2 . 17 Any 4. Typical values of this air temperature gradient are 0. These values have been standardized over the years and are commonly used in all design heating and cooling load calculations. .00 0.32 1. this gradient is incorporated into another design heat loss (HLGG) to illustrate what effect it has on the results. .76 1.76 0.37 0.5 to 2.224. .59 1.55 h0 ~R .91 0.61 0.5-mphWmd (for summer) R h| R 1. For the ASHRAE standard design heat loss calculations (HLD) made later. R 0.60 0. Also shown as Table 5 are common emissivity values for building materials. Table 4. Conductances are for surfaces of the stated emittanee facing virtual blaciibody surroundings at the same temperature as the ambient air. it is assumed that there is no air temperature gradient.08 0.10 1. .5 in. STILL AIR Horizontal . No air space value exists for any surface facing an air space of less than 0. . However. Downward MOVING AIR (Any Position) 15-mphWind (for winter) 7. Surface Conductances (Btu/hr ft^ F) and Resistances for Air [1]. Floor Convection Coefficients.14 1. floors. . and ceiling that contain a contribution from radiation as well as convection. °For ventilated attics or spaces above ceilings under summer conditions (heat flow down) see Table 4.70 h0 R h0 R Any 6.22 0.92 0.63 1.32 0.90 h.88 0. Upward Sloping—45 deg Upward Vertical Horizontal Sloping—45 deg Downward Horizontal . Values are based on a surface-air temperature difference of 10 deg F and for surface temperature of 70 F.0 F per foot.05 € = 0.Reflective Reflective reflective t = 0. Surface Emittanee Position of Surface Direction of Heat Flow Non.__ a N o surface has both an air space resistance value and a surface resistance value.37 1. d See Fig.25 h. 5. it is important to realize that these standard coefficients may no longer apply due to higher surface temperatures.62 0.Wall.73 1.68 0. The U-factors indicated in the transmission loss component include convection on the inside walls.loss due to higher air temperatures at the ceiling which can cause higher transmission and infiltration losses. .20 t = 0.45 2. For the standard ASHRAE design heat loss procedure (HLD) the convection coefficients in Reference [1] were used and these are given in Table 4.1.67 2. -46- .46 1.00 0.60 1.70 0. When radiant systems are considered. . 2 for additional data.35 1.74 0.3 . Ceiling. . 06 75to84 70to80 30 to 70 0.20 0. a schematic of the room configuration used for the calcula-.82 0.Table 5.galvanized. and a ceiling. nonmetallic paints Regular glass One surface emittance v.90 0.05 0. 5. masonry. consequences of higher mean radiant temperatures and lower air temperatures.24 0. and to investigate the effect of changing specific parameters in the design process for radiant units. For the floor. The U-factors for the transmission losses also contain outside convection coefficients for the walls and floors "and ceilings if appropriate.50 0. These include effect of higher room surface temperatures. Aluminum paint Building materials: wood.25 0. Effective Emittance E of Air Space Surface Reflectivity.. tion of the radiant design heat loss values is shown. Average in Percent Emittance t Aluminum foil.05 0. In Figure 17. convection coefficient (hc) and a U-factor (U) are specified. four walls. to be able to compare this design load with the ASHRAE standard procedure (HID) described above.11 0.84 0.12 0. an emissisivity (e). a floor.12 0.2 .Development of Design Heat Loss Procedure for Radiant Systems It is necessary with radiant types of systems to be able to estimate the design heat loss value so that units can be sized and located. It is also important for this study.72 '. Reflectivity and Emittance Values of Various Surfaces [1].bright. The ceiling is composed of several portions: heating or cooling panels and ceiling. paper. bright Aluminum sheet Aluminum coated paper.35 5 to 15 5 to 15 0.15 0. -47- . There are six surfaces specified.77 0. and changes in the infiltration heat loss term. the other <U|0 Both surfaces emitiancest 92 to 97 80to95 0..47 0.20 6. polished Steel.82 0.03 0. PANEL Heat Input £ h Wall 2 Loss-w3 GLASS I Loss-w2 •p- oo l £.h .F FIGURE 17.U c Wall 1 e.U c Loss-wl 'infiltration Floor e.h .U c Loss . h . SCHEMATIC OF ROOM CONFIGURATION USED FOR THE CALCULATION OF RADIANT DESIGN HEAT LOSS VALUES . From this information the angle factors between all of the room surfaces can then be calculated.001 Btu/hr ft^ F) value. For this analysis an air change per hour (ACH) was specified as input information and then with the room volume known an air volume could be calculated. The heating panels will be supplying heat by radiation and convection to the other surfaces and the room air. and a U-factor (U) specified. the room height. The remainder of ceiling (See Figure 17) has an emissivity (e). Also. By specifying the size of the room (length and width). the infiltration air will be affecting the overall heat balance of the room air. convection coefficient (h c ). As indicated in Figure 17 there is a contribution to the total design heat loss by the infiltration term. The four walls can be individually described by giving an emissivity (e). In addition. the contribution of glass in the outside walls can be varied by specifying appropriate values of the wall U-factor. the outside walls will transfer heat to the surroundings as will the floor and the ceiling. convection coefficient (hc) and a U-factor (U) for each wall. The following sections contain a description of the system of equations solved using a computer program and how these equations were formulated. This allows the walls to be outside or inside walls by using the actual U or a small U (0. There is no U-factor specified for the heating/cooling panels since this would vary considerably from unit to unit and it can be taken into account in the design process. At the same time. It is important to understand that all of the surfaces shown in Figure 17 are coupled thermally through their radiant exchange and their convective exchange with the room air. To -49- .For all of the equally sized panels an emissivity (e) and convection coefficient (hc) must be specified. the number of ceiling panels and their coordinate locations the geometry of room is defined. Heat Balance on Room Surfaces TA1. 5.1 . The sum of these two heat flows. algorithm based on Newton's Method. it was found that the following system of equations needed to be solved: i) Heat balance on the room surfaces (six surfaces). ii) Heat balance on the complete room. under steady state conditions. the design heat losses and other parameters were evaluated. Surface Heat Exchange Model in convective exchange with the air in the room. This results in ten equations to be solved. Each room surface area A^ as illustrated in Figure 18 is in radiant exchange with all the other surfaces and is Surface i Figure 18. iv) The definition of mean radiant temperature. q r and q c v will. This required solving a system The solution was done using an Once this system of equations was solved and all of the temperatures known.achieve this aim.2. where nine of the equations are coupled and eight of the nine are non-linear. 1r + <Icv + qcd = ° -50- (19> . be equal to the conductive heat flow through the surface as shown below. of nine non-linear equations simultaneously. iii) The comfort equations (two equations). The reason for this is that -51- .2 . A^ and per unit time. q cv £ <= convective heat transfer from surface i to air per unit area Aj h c ^ = the appropriate convection heat transfer coefficient T a = air temperature T^ = surface A^ temperature The h c ^ heat transfer coefficients selected were different for the nonradiant heating and radiant heating cases.1. qr j_ = net radiant heat transferred from surface i per unit area.i = h c > i (T i " Ta> <21> where. 5. The angle factors were calculated from algorithms available in References 5.qr. a = Stefan-Boltzman Constant FA^-Aj =• angle factor from surface i to surface j. Radiant Exchange Rate. 6. For emittances of surfaces at or above 0.2.1 . and 16.1.S 4 ei a Ti F A _A (20) where.2.q cv Convective Heat Transfer. surface reflections can be ignored from surfaces and the radiant exchange can be expressed as: q r i 4 = £i<7 Ti . This term is evaluated from the following equation. T^ «=» absolute temperature of surface A^ e^ = emittance of surface i. qcv.wherej q r = net radiant heat transfer from Aj q cv — convection between air and surface A^ qC(j = conduction through surface A^ 5.9. At = average surface to average air temperature difference. C H = height of room. 18].0. upward heat flow [1. higher air velocities and lower surface temperatures are expected.712 Btu/hr ft2 F Ceiling/floor.03 (—) 0 .041 De0-25 where. w/m2 C H where. 18].68 w/m2 C = 1 Btu/hr ft 2 F) For the radiant heating systems the following convection coefficients were used.29 H0. a) Heated Ceiling Heated Ceiling Panels [16. (At)0-25 . For the non-radiant heating calculation the following coefficients were used. downward heat flow [1.18]. (At) 0 .162 Btu/hr ft2 F Walls [17] At h c = 2. 16] h c = 0. -52- .162 Btu/hr ft 2 F Walls [16. Btu/hr ft 2 F h„ = 0. D e =• equivalent diameter (4 times the area divided by perimeter) Unheated Ceiling Portion [16. m (5. 16] h c =0.in the non-radiant heating systems. h c .2 2 . 18].3 2 h c t= 0. Ceiling/floor.05 Floor [16. 18] AtO-31 h c . C^ = overall wall conductance from inside surface to outside air *1 + Kl — l + . equivalent diameter (4 times the area divided by perimeter) Unheated Floor/Ceiling Area [16. x^ = thickness of each homogeneous section of the wall k^ = thermal conductivity of the material a^ =• conductance of each air space in the wall h Q = coefficient of heat transfer by convection and radiation at the outside surface of the wall Tj_ = inside surface temperature of surface i. + a where.3 .2.39 D. 0. Conductive Heat Transfer.18]..0.i .h c .712 Btu/hr ft 2 F b) Heated Floor Heated Floor Portion [16. h c = 0.qCc[. the heat conduction per unit area Aj_ is given by qcd.162 Btu/hr ft2oF.29(At)0-32 hc H0.712 Btu/hr ft 2 oF. T Q = outside ambient design temperature For calculation in the program the following was used to calculate Cj.T Q ) (22) where. -53- . downward heat flow Walls [16.1.18] 0.05 5. upward heat flow h c = 0.0.. Under steady state conditions.^ (Ti .08 where. I l l (23) Ci Ui ht where. Eq. Uj = The overall heat transfer coefficient from the inside air to the outside air using standard or typical ASHRAE values [1] h^ •= Convective heat transfer coefficient from inside air to inside surface i.Total Heat Loss -= 0 (24) where. This is given by the following equation. This was the typical design value for this coefficient as given in Table 2.Q i n p u t + Qpeople + Qlights Total Heat Loss .QTransmission Loss + Înfiltration Loss and. -54- .Heat Balance on the Complete Room. Total Heat Gain . These are standard or typical values used by designers and includes convection and radiation heat transfer. It is necessary from the first law of thermodynamics to maintain a heat balance on the air within the room (see Figure 17).2 . Total Heat Gain . 5. It was necessary to use only the true convection coefficient since the procedure in the calculation method accounted for the radiation. Qpeople = Internal sensible heat gain from people in space (this was set equal to zero for the design heating case). Qinput "* Heat input by supply air in the convective heating case or by panels in the panel heating case.2. 23 was used in order to eliminate the standard dual convection coefficient which includes both radiation and convection terms. In order to do this a set of comfort criteria needed to be selected.) 0 = f(Ta.Comfort Equations.1.Qliehts = Internal heat gain due to lights (this was set equal to zero for the design heating case). t. For this study. înfiltration loss = Heat loss due to infiltration air.. 5. These are. • Activity level (internal heat production in units of MET) • Thermal Resistance of Clothing (in units of clo) • Air Temperature ( F) • Mean Radiant Temperature ( F) • Humidity Level (in terms of relative humidity) • Relative Air Velocity (m/s) The comfort equations can be expressed as M . A Du where. Fanger considers the simultaneous influence of six operating variables for comfort. the Fanger Comfort Criteria [4] were chosen and were previously discussed in Section 4. Pa.3 .. Pa.6. tnrt. r.2.= f ( F p i for i = 1. m^ M = metabolic rate of person. ^transmission loss = ^ u m °^ t n e conduction losses through each of the six surfaces. MET P a = partial pressure of water vapor in the room air (a function of air temperature and relative humidity) -55- (27) . Ap u = Dubois surface area of a person. The objective of the heating or cooling system is to provide thermal comfort for people in the room illustrated in Figure 17.f( A (25) Du t^.1. T a ) t c l .. t t for i= 1 to 6) (26) Mi . t cl . 4.4 . 12 • Conduction Design Heat Loss with Room Air Temperature Gradient (HLCG) .Eqn. the functional relationship indicated by Eq. 10 • Actual Design Heat Loss (HLA) . 27 is that given by Eq. • ASHRAE Standard Heat Loss (HID) . in the case of panel heating.Eqn. For details of the development of these equations one should see Fanger 5. 13 • Actual Heat Input In the case of convective heating systems. The functional relationship indicated by Eq.T a — room air temperature. 26 is that given as Eq. The following other significant parameters were evaluated from the computer results. And finally. But.2. the actual heat input term assumes there is no heat loss from the reverse sides of the panels. 25 is that given as Eq. 14. HLCG assumes the same conductive resistance at the back of the heated panels as in the rest of the -56- . 15. the actual heat input is the same as the conduction design heat loss with room air temperature gradient.Eqn.Eqn.Other Parameters Evaluated. 11 • Conduction Design Heat Loss (HLC) . C F p i «= angle factor from the center of a seated person at center of room to surface i tj_ = surface i temperature r\ = mechanical efficiency of person tc2 = temperature of clothing surface t m r t = mean radiant temperature at center of room The functional relationship indicated by Eq. HLCG. The difference in percent between HLC and HLD. Btu/hr ft2 °F Panel Area (T . • Percentage Difference 4 .Eqn. This is necessary since each manufacturer will have different types of insulating schemes.Eqn.The difference in percent between HLCG and HLD. • Percentage Difference 1 .Eqn. Qinput = Actual Heat Input Tp = panel surface temperature T a = room air temperature • Parameter 3 .surface and Is therefore a greater value than the actual heat input term. • Operative Temperature . • Percentage Difference 2 . 9 • Parameter 1 - Qinput 0 .The difference in percent between Actual Heat Input and HLD.T a ) P where.The difference in percent between HLA and HLD. • Percentage Difference 3 . 7 • Effective Radiant Flux .Dimensionless -57- (28) . The design procedure for each type of system will then take into account the losses from the backs of the heating panels. 8 • Average Unheated Surface Temperature (AUST) . Qinput (29) Panel Area [T4 .(AUST)4] a P where. a = Stefan Boltzman Constant • Percentage Radiation - QRP (30) QRP + QCVP -58- . 1 and half glass with U = 0.Comparison of Calculated Design Radiant Loads With the Standard ASHRAE Design Load Calculation Many cases have been run for both forced air and radiant systems in order to determine the effect of various parameters and variables on the design heat loss.58.0. Walls 2. Glass distributed uniformly over the wall.half wall with U = 0.4 .3. QRP ~* radiant heat output by panels Q CV p = convective heat output by panels 5.07 Ceiling . Width = 30 ft.3 . A base configuration was selected and this was used to make initial calculations and then changes in the parameters were made in order to test their effect on the value of the design heat loss. U-Factors (Btu/hr ft 2 F): Wall 1 .U . Outside Design Temperature = 3 F Room Dimensions: Length = 30 ft.9 Floor: 0.9 Ceiling: 0.0.9 Walls: 0.where. The configuration was the following.1 Floor .9 Convection Coefficients (Btu/hr ft2 °F): -59- .U = 0.U . Height = 9 ft.07 Emissivities: Panels: 0. floors. ft. and ceiling were not changed during the operation of the system. there was no supply air.2 Comfort Variables: Metabolic Rate: . This allowed the design heat loss values to be calculated and compared with the standard ASHRAE procedure. This is due to increased infiltration as well as the increase in wall and -60- . and the room air temperature gradient was set at 0 F per foot. the air temperature gradient was set at 0. the supply air flow rate was set at 0. For radiant heating.4 -Test Case Calculations.2. For cooling cases. In order to be able to evaluate the performance of the computational scheme.75 clo (fci —1.2. the forced air heating case was taken as a test case.75 k cal/hr m 2 Clothing Resistance =0. The values given in Section 5.75 CFM per sq. 5. this would not be the case. In these calculations the convection coefficient on the walls.15 m/s Relative Humidity = 30% Infiltration Rate =0. of floor area.1. For these design calculations.See Section 5.5 air changes per hour For convection heating.75 F per foot with a reference height of 5 feet from the floor. and no lighting load was considered.1. For convection heating.1) Relative Air Velocity =0.2 were used and remained constant (except for the walls where they were a function of the At). The standard forced air heating cases are given in table 6 for various heights of the room. the number of people was set at zero. the ASHRAE Design Heat Loss (HLD) increases and correspondingly so does the supply air temperature. As the room height increases. 4 -8. DEG. F SUPPLY AIR TEMPERATURE.0 ACTUAL DESIGN HEAT LOSS.2 -8.1 61.6 26747.5 -5.2 61.9 78.4 44389.8 -4. F 77.6 30039.2 27792. .6 40482. F .2 69.3 62. F 61.7 -13.9 -7.0 15.8 62055.6 79.0 48384.4 105.8 101.8 24645. F 62.9 26443.4 56334.9 60.3 OPERATIVE TEMPERATURE.6 26747.0 12. FT PERCENTAGE DIFFERENCE 1 CONDUCTION DESIGN HEAT LOSS 1.0 68.2 60.6 62.5 61.4 59.2 22942.9 -6. BTU/HR.8 51256.4 22942.9 61.1 60.5 -5. F 69.7 46212. SQ FT A.4 109.0 25.TABLE 6.3 79.3 -8.4 -10.1 -6.5 61.6 24864.6 22796.7 116.9 -8.8 68.5 -7.5 5.STANDARD CASE CALCULATIONS 8.9 4.4 61.9 6. DEG.U.4 30549.4 36310.0 ASHRAE DESIGN HEAT LOSS.S.1 ROOM AIR TEMPERATURE.6 24864.0 126.4 30549.8 29895. BTU/HR PERCENTAGE DIFFERENCE 2 CONDUCTION DESIGN HEAT LOSS 2.0 60.1 103.5 -12. DEG. BTU/HR EFFECTIVE RADIANT FIELD.4 62.6 68.3 -11. BTU/HR 24796.4 56334.1 -12.5 -7.6 34124.1 69.5 -3.5 77.1 -6.9 -14.7 61.0 9. BTU/HR 25642.5 -3.0 32659.0 -8.9 PERCENTAGE DIFFERENCE 4 -7.8 137.1 -7.9 -6.9 53297.0 59.0 58212.7 ROOM HEIGHT. DEG.7 46212. DEG.4 36310.T.7 -11. DEG. BTU/HR PERCENTAGE DIFFERENCE 3 ACTUAL HEAT INPUT.8 -4.2 78.1 3.8 MEAN RADIANT TEMPERATURE.8 77.4 3. FORCED AIR HEATING .2 38556.7 59.1 69.8 26762.1 35399.0 10.3 59.0 20.2 FLOOR TEMPERATURE.1 4.4 28728.7 62.0 5.0 -8. These higher air temperatures are consistent with the results presented in Table 1 for comfort conditions. It is also important to notice that the room air temperature for comfort is about 77 F for the 8 feet high and almost 80 F for the 25 feet high room. overestimates the calculated heat loss HLC or HLCG by about 7% for an 8 feet high room and by about 3% for a 25 feet high room even with a temperature gradient. Tables 10 and 11 give the results for the forced air heating system standard case with different infiltration rates and for a 15 ft. Similar results are exhibited except that the ASHRAE design heat loss. underestimates the heat loss by about 2% for the 25 feet high room with a temperature gradient of 1.75 CFM/ft .glass areas as the wall height is raised causing larger heat losses. Tables 7. This shows up also in a reduced value of AUST with increasing height. high room respectively. This is due to the mean radiant temperature dropping because of more glass surface in the higher room and therefore a higher air temperature being required to satisfy the comfort equations. and 12 that the supply air temperatures are not appropriate. this -62- . HLD.5.0 and 1. and 25 ft.5 F per foot. the infiltration air leaving the room at the ceiling level is at a higher temperature due to an air temperature gradient. It should be noted in Tables 10. The air flow rate was set at 0. As the room height increases. This underestimation can be up to 16% at 4 air changes per hour.5 F per foot respectively. and for higher heat losses as found here. 1. The ASHRAE design heat loss. HLD. 11. The results in Table 12 are for the same conditions as in Table 11 except that the tJ-factors were increased to what might be expected in industrial situations. 8 and 9 show similar results as Table 6 except that the air temperature gradient was changed to 0. Comparison of these results show that the ASHRAE design heat loss calculation can underestimate the size of the heating load for high (greater than 2) infiltration air changes. 6 79.3 105.1 69.0 58212.4 -7.2 FLOOR TEMPERATURE.9 45604.4 -10.0 5. BTU/HR PERCENTAGE DIFFERENCE 2 r CONDUCTION DESIGN HEAT LOSS 2 .5 115.7 -5.TABLE 7.8 101.0 -8.7 -7.6 68. F 69.7 -7.0 24791. 5 ° F / F T .4 -7.7 -11. DEG.2 136.0 32659.0 103.3 36006. SQ FT A .6 40482.3 36006.3 OPERATIVE TEMPERATURE.1 60.1 3. F SUPPLY AIR TEMPERATURE.0 ROOM HEIGHT.7 PERCENTAGE DIFFERENCE 3 ACTUAL HEAT INPUT.1 -7.7 61.9 26443.0 -6.4 28728. F 61.0 60. BTU/HR 22894.9 -8.9 55322.0 25.4 44389.6 34124.9 4.3 79.3 -8.4 61.9 -14.4 22894.8 68.6 22796.4 30379.3 62.8 77.7 62.5 77.9 78.9 53297.2 27792.3 -11. BTU/HR EFFECTIVE RADIANT FIELD.1 -12.2 61.7 -5.6 62.7 59.WITH GRADIENT = 0 .7 -7.4 62.1 4. DEG.0 24791.9 45604.4 3.6 30039.1 35399.2 60.0 9.4 -7. T .8 29895.3 59.8 62055.0 10. DEG.0 30379.8 MEAN RADIANT TEMPERATURE.1 . 8.3 109.0 ACTUAL DESIGN HEAT LOSS.0 ASHRAE DESIGN HEAT LOSS. BTU/HR 25642. FORCED AIR HEATING .2 -8.0 -8.9 61. F 62.0 68.1 69.9 6. DEG.8 24645.5 61.5 61.0 15.4 59.5 -12.0 12.0 20.1 61.9 55322.4 -8. U .1 ROOM AIR TEMPERATURE.2 78. DEG.0 59.6 -5.7 PERCENTAGE DIFFERENCE 4 -7.0 48384. FT PERCENTAGE DIFFERENCE 1 CONDUCTION DESIGN HEAT LOSS 1 .6 126.8 51256.9 60. S .2 26646. BTU/HR. DEG.2 69.0 -6.2 38556.6 -5.1 -7. BTU/HR 24796.5 5.7 -13. F 26646.8 26762. F 77. 0 -3.2 -1.4 -7.6 62. BTU/HR.6 30039.2 60.8 -6.1 -7. DEG.1 69.4 62.7 -13.4 44389.3 59.0 25.8 68.4 22991. ROOM HEIGHT.WITH GRADIENT = 1.8 51256.5 5.0 ACTUAL DESIGN HEAT LOSS. BTU/HR EFFECTIVE RADIANT FIELD.4 138.9 78.0 -8.7 59.4 46819. DEG.2 -1.1 60.9 -5.8 101.8 24645.3 -6.9 60.8 PERCENTAGE DIFFERENCE 1 CONDUCTION DESIGN HEAT LOSS 1.2 69.9 26443.2 24937.1 ROOM AIR TEMPERATURE.8 77.8 62055-1 3.0 ASHRAE DESIGN HEAT LOSS.S. F 62. DEG.0 10.5 FLOOR TEMPERATURE.3 -8.9 -5.3 127. SQ FT A.4 28728.5 36614.0 20. FT 8.4 59.6 30719.8 29895.0 68.1 61.6 40482.TABLE 8.5 -12.2 61. F 61.T.1 -12.1 4.0 48384.0 59.0 58212. BTU/HR PERCENTAGE DIFFERENCE 3 ACTUAL HEAT INPUT.0 60.5 61.7 61.0 12.4 PERCENTAGE DIFFERENCE 4 -7.8 MEAN RADIANT TEMPERATURE.6 68.5 26848.U.3 62.1 35399.2 38556.2 24937. F 77. DEG.1 69. F .9 4. BTU/HR PERCENTAGE DIFFERENCE 2 CONDUCTION DESIGN HEAT LOSS 2.3 OPERATIVE TEMPERATURE.9 6.2 -8. FORCED AIR HEATING .4 -8. DEG.6 22796.6 79.7 62.5 61.0 -8.4 3.4 61.2 27792. BTU/HR 25642.8 116. DEG.9 53297. BTU/HR 24796.5 22991.0 15.9 57347.3 -6.5 109.9 57347.9 61.0°F/FT.5 26848.9 -14.8 26762.5 -5.4 -10.5 77. F 69.1 103.6 30719.5 -5.3 79.8 -6.3 -11.0 9.4 46819.0 -3.0 32659.9 -8.2 78.5 36614. F SUPPLY AIR TEMPERATURE.0 5.7 -11.6 34124.4 105. 1 -12.9 6.3 -5.1 69.1 61. F SUPPLY AIR TEMPERATURE.0 10.7 61.3 -8.T.8 51256. DEG.1 31059.6 30039.9 78.3 59.0 48384.9 -6.0 59.9 59372.8 77.2 38556.1 -7.4 28728.9 -14.1 4. F 61.9 -3.9 53297.9 -8.0 68.9 128.4 59.0 60.0 9.8 68.9 4.4 23088.3 62. BTU/HR 25642.6 22796.9 48034.2 116.5 -0.9 26443.4 PERCENTAGE DIFFERENCE 4 -6.3 27051.3 OPERATIVE TEMPERATURE.0 12.6 62.8 26762.7 2.7 37221. F 62. DEG.8 ROOM HEIGHT.0 -8.6 40482.0 32659. BTU/HR EFFECTIVE RADIANT FIELD.4 -6.0 ASHRAE DESIGN HEAT LOSS. BTU/HR 24796.3 27051.0 23088.9 -3.6 68.6 79.0 -8.0 5. FORCED AIR HEATING .9 59372.8 24645.8 MEAN RADIANT TEMPERATURE. BTU/HR PERCENTAGE DIFFERENCE 2 CONDUCTION DESIGN HEAT LOSS 2.3 79. DEG.7 37221.9 -6.7 -11.5 5.S.1 ROOM AIR TEMPERATURE. DEG.5 61. SQ FT A.2 61.8 62055.2 103.8 -4.3 -11.0 25.9 60. BTU/HR PERCENTAGE DIFFERENCE 3 ACTUAL HEAT INPUT. FT PERCENTAGE DIFFERENCE 1 CONDUCTION DESIGN HEAT LOSS 1.3 -5.7 140.WITH GRADIENT = 1.2 60.0 58212.1 3. F 77.4 62.4 44389.0 15.TABLE 9. DEG.5 -12.6 34124.5 77.5°F/FT.4 25083.0 20.7 62. F .2 69.7 2.0 FLOOR TEMPERATURE.4 3.0 ACTUAL DESIGN HEAT LOSS.4 -8.5 -0. DEG.4 25083. F 69.6 105.8 -4. BTU/HR.1 35399.U.8 29895.1 69. 8.7 59.8 110.8 101.2 27792.2 -8.9 48034.4 6-1.1 3!1059.2 78.1 60.5 61.9 61.7 -13.4 -10. 1 61. F SUPPLY AIR TEMPERATURE.TABLE 10.3 60.1 66599.1 61.6 i« ACTUAL HEAT INPUT.9 5.9 68038.8 68.0 64800.7 -12.6 104778.6 78.2 -0.0 5.9 188. F 61. DEG. BTU/HR.U.1 46407.S.0 44584.0 61.6 78. SQ FT A. BTU/HR 47304.6 78.8 -12. DEG.6 5.0 4.6 MEAN RADIANT TEMPERATURE.0 OPERATIVE TEMPERATURE.0 3. BTU/HR PERCENTAGE DIFFERENCE 3 86792.0 61.3 60. F 60.7 106985.9 -1.0 2. BTU/HR 49667.9 *'2.6 62954. F 68.7 106985.3 147.T. AC/H HEIGHT AND VARIABLE INFILTRATION RATES 1.8 -1. F 78. F .7 EFFECTIVE RADIANT FIELD.8 68.3 86408.0 99792. DEG.0 61.8 46407.2 99695..STANDARD CASE-15 FT AIR CHANGES PER HOUR.8 68. DEG.7 -12.0 5.1 66599.0 82296. DEG.8 5.0 ACTUAL DESIGN HEAT LOSS.3 ROOM AIR TEMPERATURE.0 5. FORCED AIR HEATING.2 FLOOR TEMPERATURE.1 167.7 61.2 86792.0 .7 -2.9 81325.9 PERCENTAGE DIFFERENCE 1 CONDUCTION DESIGN HEAT LOSS 1. BTU/HR PERCENTAGE DIFFERENCE 2 CONDUCTION DESIGN HEAT LOSS 2.3 60.9 2. BTU/HR 1 PERCENTAGE DIFFERENCE 4 -1. DEG.1 61.1 126.7 -12.5 7. ASHRAE DESIGN HEAT LOSS.6 -5.5 7. 1 59.8 79.0 ACTUAL DESIGN HEAT LOSS.3 -2.8 156.7 -14. BTU/HR PERCENTAGE DIFFERENCE 1 ON HEIGHT AND VARIABLE INFILTRATION RATES ACTUAL HEAT INPUT.4 68.4 68.9 9.8 59.9 13.1 1.0 3.6 6.2 186395.6 99925.0 160272.1 149235. BTU/HR 77597. SQ FT A.3 59.8 59.0 101952. DEG.U.S.0 -0.8 195.8 MEAN RADIANT TEMPERATURE. BTU/HR PERCENTAGE DIFFERENCE 2 CONDUCTION DESIGN HEAT LOSS 2.3 2.-S:. F 79.7 -14. F SUPPLY AIR TEMPERATURE.0 72792. F 59.1 233. DEG.T.1 131010. AC/H 2.8 16.6 68840.8 79.3 59.0 112075.9 13. DEG. DEG.0 112075.6 108682.4 68.6 6.7 -14.3 OPERATIVE TEMPERATURE.1 74915.8 79.1 59.1 59.STANDARD CASE WITH 25 FT.4 6. DEG.7 139767.9 9.3 271. BTU/HR.3 59.0 ASHRAE DESIGN HEAT LOSS.«< FORCED AIR HEATING .0 4.3 74915.3 2.0 131112. &*':••:~? FB-Hras p».7 59. DEG. BTU/HR PERCENTAGE DIFFERENCE 3 I 1.0 -5. F 59. INFILTRATION.2 186395.8 170852.9 CONDUCTION DESIGN HEAT LOSS 1.2 162095. F 68.4 -14.5 PERCENTAGE DIFFERENCE 4 EFFECTIVE RADIANT FIELD.8 16. BTU/HR 6.3 FLOOR TEMPERATURE.•#£e gXrv-^gj TABLE 1 1 .1 149235.1 ROOM AIR TEMPERATURE. F .8 59. 3 47.TABLE 12.7 217. DEG.8 18. F 65.5 42.lc .6 166345. BTU/HR 18.S.0 -6.8 6.» .5 42.7 0. DEG.9 PERCENTAGE DIFFERENCE 1 CONDUCTION DESIGN HEAT LOSS 1.. -2.0 3.3 -30.4 230620.5 ROOM AIR TEMPERATURE. ceiling ? U = 1 . walls. DEG.0 ASHRAE DESIGN HEAT LOSS.0 223478.5.8 207062.8 18.4 EFFECTIVE RADIANT FIELD. F 88.5 88. BTU/HR 161338.T.0 ACTUAL DESIGN HEAT LOSS.9 OPERATIVE TEMPERATURE. Btu hr ft F 25 ft glass ° / .5 88.7 47. DEG.6 154195.2 -12. F SUPPLY AIR TEMPERATURE.2 PERCENTAGE DIFFERENCE 4 -7.5 0. F .9 45.2 195979.3 -30.8 259.HEIGHT AND VARIABLE INFILTRATION INFILTRATION.5 MEAN RADIANT TEMPERATURE.1 125629. DEG.5 42. DEG.8 188837.U. BTU/HR PERCENTAGE DIFFERENCE 2 CONDUCTION DESIGN HEAT LOSS 2.1 65.8 18. F 45.25 Btu/hr ft 2 °F.1 65.3 -30.0 4. SQ FT A.6 343.9 45.. BTU/HR 135810.0 164970.0 2.0 223290.1 -30.5 0.0 194130.1 65. BTU/HR PERCENTAGE DIFFERENCE 3 ACTUAL HEAT INPUT.6 166345.STANDARD CASE WITH U ..7 265261. FORCED AIR HEATING .7 11.0 247778. „„ „„.5 88.floors.7 11. F 42.9 45.0 125629.„„ = 0.8 6.7 47.0 247778.0 FLOOR TEMPERATURE. AC/H 1.7 301.8 207062.8 119554.7 47. BTU/HR.2 -7. In this calculation. In this procedure the panel temperature was assumed as input information and a trial and error procedure was used to determine the required area for the heat loss from the space. The same base case was taken as in the forced air system except a single radiant heating panel was used to supply heat to the room and there was no heated supply air. ft.9 and its convection coefficient was as previously specified. and the calculated area was 176 sq. The convective calculations appear to be reasonable and correct and do not show any unusual results.Radiant Panel Heating Systems Calculations 5. At 120 F one manufacturer's procedure indicated 453 sq. They indicate that the program is calculating values that are expected and show that the ASHRAE standard design procedure tends to slightly overestimate design losses even with an air temperature gradient present except for high (above 2) air changes per hour of infiltration. In Table 13. the area required for heating with panels reduced as the panel temperature increased. ft. This information appears to verify the -69- . ft. and the other manufacturer's procedure indicated 415 sq.Single Panel Radiant Heating Cases. For 120 F. The areas calculated here were compared with the required area from two manufacturers of hydronic panels and showed quite close agreement. As expected. The calculation here indicated 439 sq.would have to be raised to approximately 3 CFM/ftr to yield reasonable supply air temperatures. At 180 F the two numbers were 216 and 185 sq.5 . ft. the results for panel surface temperatures from 120 to 180 F are shown for the base case room.1 . 5. This calculation does not affect the design load calculations . approximately 49% of the ceiling area was covered with radiant panels while for a 180 F panel temperature approximately 20% of the ceiling was covered with radiant panels. the emissivity of the panel heater was set at 0.5. ft. TABLE 13.2 25644.6 109.2 25644.4 7. BTU/HR PERCENTAGE DIFFERENCE 2 CONDUCTION DESIGN HEAT LOSS 2.8 74.1730 1.3 -8.0 180. DEG.7 66.0 OPERATIVE TEMPERATURE. 5 68.1 24005.8 74.2 -4.2 24836. F 77.0 160.4 72. SQ FT A.7 302.0 170.2 -4.F 1.5 7. BTU/HR 23664.2111 1.2497 PARAMETER 3.6 25646. F .2 PERCENTAGE RADIATION 95.SQ FT 53.4 72.5 25662. BTU/HR PERCENTAGE DIFFERENCE 3 ACTUAL HEAT INPUT. PANEL TEMPERATURE.2 -4.0937 1.7 25.6 125.4 26762.4 ACTUAL DESIGN HEAT LOSS.6 -7.2 23655.7 66.0 PANEL AREA REQUIRED .3 95.4 68.7 66.8 40.0 77.4 26762.6 24613. DEG.5 -11.5 7.7 80.4 72.0091 1.8 -11.3 94.4 95.5 -11.8 23667.0089 0.3 95.6 68. DEG.0 22.4 26762..0 141.2 .3 95.3 23678.T. BTU/HR.6 -11.0124 0.SQ FT.5 7.1338 1. DIMENSIONLESS 0.9 74.1 23674.6 -11.1 -4.1 25645.1 -4.6 -11.0168 0.1 23676.2 -4.6 -8.0 33.7 68.4 PERCENT CEILING COVERED BY PANELS 48.1 25647.4 25650.5 HEAT OUTPUT PER UNIT PANEL AREA. BASE CASE FOR RADIANT PANEL HEATING: PANEL HEATING .1 -4.0142 0. -4.0 19.4 26762.CEILING HEIGHT = 9 FT.0523 1.4 25650.4 PARAMETER 1.0 77.5 68.2 25662.4 72. BTU/HR.6 7.0 140.0098 0.4 -4.8 359.1 -11. BTU/HR.6 ASHRAE DESIGN HEAT LOSS.0 77.7 25654.0 -7.9 74.6 -10. SQ FT 438.6 25646.1 -4.4 26762.5 7.7 23671.1 25647.0081 PERCENTAGE DIFFERENCE 1 CONDUCTION DESIGN HEAT LOSS 1.0109 0.7 66.6 28.4 24459.4 26762.0 77.7 66.7 197.5 FLOOR TEMPERATURE.9 175.0 23679.2 -4.3 -9.4 72.7 25654.4 68.0 130.0 77. DEG.3 95.7 24735.0 77.S. DEG. F 72.7 74.8 24263.7 66. F 74.4 74.5 -11.2 -4.U. F 66.4 -4. BTU/HR 26762. D E C F 120.7 MEAN RADIANT TEMPERATURE.1 25645.2 258. BTU/HR PERCENTAGE DIFFERENCE 4 EFFECTIVE RADIANT FIELD.2 -4.5 7.7 224.0 150.9 ROOM AIR TEMPERATURE.4 95.4 72.5 68.9 66. The room air temperature for comfort conditions in the radiant cases (about 67 F) is 10 F less than in the forced air case which reduces infiltration loss. This result is not significant in light of the many assumptions made in both cases. keep in mind that comfort conditions were satisfied at the center location for a seated person and that due to radiant temperature asymmetry discomfort could be experienced at the higher panel temperatures. The values for floor temperature. operative temperature. this reduction in infiltration loss does not overcome the increased loss due to higher surface temperatures. room air temperature. floor. This is significant since it illustrates that the radiant systems heat surfaces which in turn heat the occupants and the air while forced air systems heat the air which then heats the occupants and the surfaces. The design heat loss calculated here (HLC) is about 4% below the ASHRAE standard design heat loss calculation (HLD).calculation procedure since reasonable agreement is found with rated heating panels. Comparison with the forced air case shows about 3% more loss in the radiant situation. However. effective radient field and AUST remain relatively constant as the panel temperature increases. and ceiling temperatures experienced in the radiant system than in the forced air systems. mean radiant temperature. causing additional heat loss through the surfaces. Parameter 1 is a "pseudo" overall heat transfer coefficient (defined by -71- . For the radiant case the AUST remained at about 68 to 69 F and in the forced air system it ranged between 60 to 62 F. Normally. the higher panel temperatures would be used in rooms with higher ceilings. It should be noted in Table 13 that higher floor temperatures are present in the radiant case than in the forced air case. Parameter 1 and Parameter 3 were calculated just to observe their behavior in the radiant types of systems. Also. This is attributed to the higher wall. 475 i I 175 120 130 140 panel FIGURE 19. 150 temperature 160 Cdeg. 170 f) REQUIRED HEATING PANEL AREA AS A FUNCTION OF PANEL TEMPERATURE 180 . 7 24342.0 68.3 -11.0 76.1 24468. DIMENSIONLESS 0.U.5 74.7 7. F .4 26762. BTU/HR EFFECTIVE RADIANT FIELD.94 316.6 66.6 -8.8 PERCENT CEILING COVERED BY PANELS 35.F 1. DEG.4 80. BTU/HR 23628.9 288.7 24585.6 95. DEG.1 FLOOR TEMPERATURE.1 -3.4 26762. BTU/HR.5 72. BTU/HR PERCENTAGE DIFFERENCE 3 ACTUAL HEAT INPUT.1 25728.2194 PARAMETER 3.0138 PANEL AREA REQUIRED . F 66.0 MEAN RADIANT TEMPERATURE.1 25728.0990 1.1 77. DEG.SQ FT 76.0392 1.5 68.1 25654. F 77.9 -3.5 25858.3 68.8 76.S.4 24203.8 ASHRAE DESIGN HEAT LOSS.1 -3.5 -11.5 23735.8 PARAMETER 1. DEG.6 25796.SQ FT.7 -11.7 66.2 33.5 25858. BTU/HR 26762.9 67.p^ m^ w^ m r-^ TABLE 14.1 PERCENTAGE RADIATION 95.3 275.0 PERCENTAGE DIFFERENCE 4 -9.6 84. EFFECTS OF PANEL EMISSIVITY WITH PANEL PANEL EMISSIVITY = 140°F 0.T.4 -4.6 .9 ROOM AIR TEMPERATURE. BTU/HR PERCENTAGE D1FFERENCE 2 CONDUCTION DESIGN HEAT LOSS 2.7 68.9 89.0 23684.88 0. -3.9 -3. BTU/HR.6 25796.4 72.3 74.4 72.0131 0.? 7.0124 0.3 95.5 7.9 -11.0 -8.6 -3.4 -4.4 ACTUAL DESIGN HEAT LOSS. DEG.6 HEAT OUTPUT PER UNIT PANEL AREA. F 72.0117 0.4 25654.1590 1.5 32.1 95.92 0. BTU/HR.8 301.0 30.6 -9. SQ FT A.7 74.4 26762. F 74.90 0.' SQ FT PERCENTAGE DIFFERENCE 1 CONDUCTION DESIGN HEAT LOSS 1.9 23782.4 7.6 OPERATIVE TEMPERATURE. 88 and 0. -74- . Table 14 shows a comparison for the different assumptions concerning the radiant panel emissivity. Some of the other quantities showed only slight changes as the emissivity was varied. Figure 19.8 and 0. 28) and Parameter 3 is a dimensionless factor related to the radiant exchange process (defined by Eqn. Manufacturers indicate that a panel emissivity of 0. shows the expected nonlinearity of the required panel area as a function of panel temperature for the radiant base case given in Table 13. In Table 15.8 to 0.9 is typical over the life of the radiant panel. a value of 0. This resulted in significant changes in required radiant panel area (a 13% drop in area as emissivity changed from 0. Tables 16. Min (18) has made this point in his work and indicated that it is a difficult parameter to evaluate because of geometrical considerations.94). The only other variable to change significantly with this change in surface emissivity was the floor temperature which went from 72 to 76 F as emissivity went from 0.95 for a situation when the panel temperature was at 140 F. This caused the required panel area to increase by only 3%. There does not appear to be any significant trends to either of these parameters. 29).88 to 0. This calculation was carried out since there is a great deal of uncertainty concerning the value of the convection coefficient from surfaces when there is a large delta T such as exists in the radiant panel case.Eqn. 17 and 18 show the effects obtained when the convection coefficient for the radiant panel is changed by a factor of 2. The emissivity was varied between 0. For the remainder of the calculations. the emissivities of the walls.94 for a panel temperature of 140 F. This variation in emissivity also affected the heat output per unit area and parameters 1 and 3.9 for the surface emissivities has been used and the calculations do not appear to be sensitive to changes in the surface emissivity. floor and ceiling were varied between 0.95. 5 and 10 respectively. 9 33. BTU/HR PERCENTAGE DIFFERENCE 2 CONDUCTION DESIGN HEAT LOSS 2.8 299.T.4 23882.3 80.5 25650.2 24645.2 24263.7 MEAN RADIANT TEMPERATURE.0 77.3 33.7 -1.1 25317.3 95.6 95.4 72. DIMENSIONLESS 0.9 HEAT OUTPUT PER UNIT PANEL AREA.1 23678.2 68. F 66. F .5 26303. F 76.3 78. BTU/HR.SQ FT.9 -4.0937 1.4 72.5 69.0 23671.15.6 33.7 66. BTU/HR 23700. DEG.4 95.4 ACTUAL DESIGN HEAT LOSS.80 0.0127 0. BTU/HR.4 26762.8 66.2 -5. BTU/HR EFFECTIVE RADIANT FIELD.6 -11.U.5 25650. EFFECTS OF WALL.7 -1.1 25984.4 26762.0124 0.9 73.9 -4.7 -2.1 -11.8 23671. DEG.3 302. BTU/HR 26762.4 26303. FLOOR AND CEILING VITY WITH PANEL TEMPERATURE = 140°F 0.SQ FT 84.1 25317.4 25015.4 7.2 304. DEG.0 77.5 7.0131 0. DEG.8 PERCENTAGE RADIATION 95.3 -10. BTU/HR.1 25984.0 PERCENTAGE DIFFERENCE 4 -6.4 -11. F 71.5 -7.0677 PARAMETER 3.9 68.90 0.5 -11.7 66.2 -5.0120 WALL EMISSIVITY PANEL AREA REQUIRED .4 7.1218 1.85 0.95 295.0 ROOM AIR TEMPERATURE.F 1.6 82.4 74.S.1 PARAMETER 1.9 77.5 7.2 PERCENT CEILING COVERED BY PANELS 32.8 ASHRAE DESIGN HEAT LOSS. F 72.4 FLOOR TEMPERATURE.9 -9. DEG. SQ FT PERCENTAGE DIFFERENCE 1 CONDUCTION DESIGN HEAT LOSS 1.1531 1.0 OPERATIVE TEMPERATURE.4 72.8 76.5 68. BTU/HR PERCENTAGE DIFFERENCE 3 ACTUAL HEAT INPUT.7 -2. SQ FT A.4 26762. 4 ACTUAL DESIGN HEAT LOSS.9 -10.4 74.5 38.3 -4.5 25618.1 67.9 25637.4 -10.9 24053.3 23719. BTU/HR PERCENTAGE DIFFERENCE 4 EFFECTIVE RADIANT FIELD.5 -8.3 72.7 ASHRAE DESIGN HEAT LOSS.0 18.0177 0.3 68.2 -4.0 -7.8 6.9 167.5 25618.2 FLOOR TEMPERATURE.F 1.3 PERCENTAGE DIFFERENCE 1 CONDUCTION DESIGN HEAT LOSS 1. SQ FT A.1 74.2 -4.0 -10.3 72.7 23841.2 -4.4 OPERATIVE TEMPERATURE.1 67.0 180.9 -10.2 68.3 72. F 76.8 24749.2 -4.1 67.4 76.2 25626.4 26762.8 23839.SQ FT 56.1133 1. F 72.4 76.0 PANEL AREA REQUIRED . F .2360 1.7 70.5 25618.8 24847.2 246.S.9 131.9 25620.1 67. BTU/HR 23827.2 68.5 74.5 -4. BTU/HR 26762.0 148.8 21.5 4.9 -10.2 91.3 -4.2 23833.6 HEAT OUTPUT PER UNI. F 67.3 72.7 343.9 -10.T.3140 PARAMETER 3.3 24297.2 114.0691 1.3 -4.4 26762.0 140.3 74.TABLE 16.8 6.0 27.2747 1.0115 0.0103 0.8 68.3 72.1 -9.5 25618.3 6.2 PARAMETER 1.4 76. DEG. BTU/HR.3 25637.6 -11. DEG.2 91.8 25618.8 23841.T PANEL AREA.3 -4. F 74. D E C F 120.1961 1.8 25618.0 150.2 25626.1 91.1 24484.3 99.SQ FT.8 6.4 76.U. EFFECTS OF CHANGING PANEL CONVECTION CIENT BY A FACTOR OF TWO PANEL TEMPERATURE.2 -8.5 ROOM AIR TEMPERATURE.0150 0.0130 0.9 25620. BTU/HR.9 25619.4 26762.1 32.8 214.8 -11.4 26762.3 PERCENT CEILING COVERED BY PANELS 46.3 -4.9 24632.4 68.8 6.4 74.3 72.4 26762.5 74.4 26762.3 -4.2 91. BTU/HR. SQ FT 418.8 6.0085 CONDUCTION DESIGN HEAT LOSS 2.2 PERCENTAGE RADIATION 91.2 68.5 76.1 288.0093 0.0 170.2 91.8 6.9 -10.4 23. DEG.1 84.1 91.0 160.5 68.4 76. DEG.1 67. BTU/HR PERCENTAGE DIFFERENCE 2 fc -4.1 MEAN RADIANT TEMPERATURE.1 67. DEG.3 -4. DIMENSIONLESS 0.9 25619. BTU/HR PERCENTAGE DIFFERENCE 3 ACTUAL HEAT INPUT.0 23839.5 23841.5 -7.0 130.4 188.1554 1. 5 -4.0 72.S.0 130.5 25563.6 73.1 24689.U.2 MEAN RADIANT TEMPERATURE.0149 0.0 5.4 217.5 -4.6 73.0 180.7 67.5 73.2 369.3 24228.0 67.7 25561.1 75.5 73.5 79.5 -4.4 26762.2 68.0 PERCENT CEILING COVERED BY PANELS 41.6 25561.0118 0.1 148.7 67.7 130.5 -9.2 68.7 67.8 24560.0 72. DEC F PANEL AREA REQUIRED .4 -9. BTU/HR PERCENTAGE DIFFERENCE 2 CONDUCTION DESIGN HEAT LOSS 2.2 68.5 25563.2 21. DIMENSIONLESS 0.5 -4. F 68.5 -9.9 81.4 26762.9 302.3 24.9 -4.9 24170. BTU/HR PERCENTAGE DIFFERENCE 3 ACTUAL HEAT INPUT.SQ FT 64.9 80.0 PERCENTAGE RADIATION 81. BTU/HR 26762. SQ FT A.4 -9.0132 0.4185 1.0204 0.8 189.1 75.4 -9.6 167.0 72.0097 CONDUCTION DESIGN HEAT LOSS 1. DEG.7 25565.1 24236. BTU/HR PERCENTAGE DIFFERENCE 4 EFFECTIVE RADIANT FIELD.5 -4.0 5.0 5.2433 1.T.7 PERCENTAGE DIFFERENCE 1 -9.7 PARAMETER 1. BTU/HR.SQ FT.1 18.3768 1.7 28.8 67.0 81.4 26762.F 1.5 -4.1 75.0 24232.4 -9.0 150. F . F 73.4 26762.0 170.9 FLOOR TEMPERATURE.4 73.0 10.9 148.1 80.6 ROOM AIR TEMPERATURE.5 -4. BTU/HR.4 26762.2 -7. DEG. SQ FT COEFFICIENTS BY A FACTOR OF FIVE 120.0 5.TABLE 17.5 HEAT OUTPUT PER UNIT PANEL AREA.0 72. BTU/HR 24233.1 75.5 25564.7 25561.4 26762.0172 0.2893 1.9 -4.0 24237.0 5.0 72.6 25562.9 112.5 25563.4 ACTUAL DESIGN HEAT LOSS.8 95. BTU/HR.5 -4.1 33.5024 PARAMETER 3.6 24795. F 72.3 73.5 -4.9 -8.3 -7.6 25561.1 24236.6 25562.1 75.5 -4.9 80.9 -10.8 ASHRAE DESIGN HEAT LOSS.7 25565. DEG.5 25564.0107 0.3337 1.7 -7.2 68.6 24235.1 OPERATIVE TEMPERATURE.0 72.7 -8.8 -9.9 80.0 5.7 67.2 68.0 166.9 254.9 24882.5 16.5 25563.5 -4.8 67.0 • 5.2 68.4606 1.0 24391. DEG. EFFECTS OF CHANGING PANEL PANEL TEMPERATURE. DEG.5 -4. F 75.1 75.0 160.5 23866. 8 -7.0117 PERCENTAGE DIFFERENCE 1 CONDUCTION DESIGN HEAT LOSS 1.9 -7.8 -7.7 -4.8065 PARAMETER 3.0159 0.5 73. SQ FT A.7 25514.2 24779. DEC F 71.2 25516.SQ FT.4 69.6 FLOOR TEMPERATURE.5 72.8 -7.6 24943.4 69.5 68.8 199.2 PARAMETER 1.7122 1.4 69. BTU/HR 26762. BTU/HR 24671. DEC F 73.7 -4.3 24678.6 256.0 170. SQ FT 313.4 69.9 24681. DIMENSIONLESS 0.0246 0.9 ASHRAE DESIGN HEAT LOSS.8 68.6 13.0 114.5 68.9 HEAT OUTPUT PER UNIT PANEL AREA.9 20.6 -4.4 72. DEC F 120.2 67.1 -9.5134 1.7 3.5 73.4 72.0 3.5 73.4 26762.0 130.7 184.7 -4. BTU/HR PERCENTAGE DIFFERENCE 3 ACTUAL HEAT INPUT.1 -8.0180 0.7 71.4 24683. BTU/HR.5 25519.7 71.4 25518.2 134.3 67.0 140.9 140.7596 1.8 95.7 25514.5 73.0 3.7 71.0 3.5 72.0128 0.7 24869.2 25516.4 -7.8 25513.3 24335.0 214.F 1.8 -7.6 25513.3 25517. BTU/HR.0 3. DEC F 69.6 24677.6 124.4 26762.6 24071.3 25517.6 24682.7 71.4 23.0 150.2 67.7 71.4 25518.0 PANEL AREA REQUIRED .5 17.TABLE 18.U.8 15.2 67.5 ROOM AIR TEMPERATURE.0 -10.4 ACTUAL DESIGN HEAT LOSS.8 -7.6 68.0 67.8 -7.6 -4.8 28.4 26762.0 3.6161 1.7 71.6 -4.5 73.SQ FT 76.4 26762.5 72.2 67.5 25519. DEC F 72.4 69.7 -4.5 73. BTU/HR PERCENTAGE DIFFERENCE 4 EFFECTIVE RADIANT FIELD. BTU/HR.6642 1.4 69.0 160.0 3. EFFECTS OF CHANGING PANEL CONVECTION CIENTS BY A FACTOR OF TEN PANEL TEMPERATURE.0208 0.1 -6.7 -4.6 PERCENT CEILING COVERED BY PANELS 34.4 -7.7 -4.2 25515.9 176.7 ' -4.4 MEAN RADIANT TEMPERATURE.8 PERCENTAGE RADIATION 69.1 159.4 68.6 -4.3 24525.0 180.T.5 OPERATIVE TEMPERATURE.9 24667.5657 1.0142 0.0 154.7 -4.5 72.4 26762.8 -7.2 25515.7 -4.S.8 24684. BTU/HR PERCENTAGE DIFFERENCE 2 CONDUCTION DESIGN HEAT LOSS 2.2 67.4 26762. DEC F .7 -4.1 68. 150. The percent radiation delivered by the panels does change as the convection coefficient is increased and this variation is illustrated in Figure 20.Effect Due to Infiltration for Radiant Panel Systems.Effect of Glass Distribution. By increasing the convection coefficient by 500% (Table 17) the area is reduced by 15. approximately one percent more loss would be added to this number (See Section 5. Different values of infiltration rates (0. Different combinations and quantities -79- . The percent difference in the design loads as a function of infiltration is shown in Figure 21.5. 5.0 air changes per hour) were assumed for the base case configuration and these results are given in Tables 19.2 . while the room air temperature and parameters 1 and 3 decrease significantly. it is not a significant variation if the convection coefficient is known within a factor of two. and 170 F panel surface temperatures respectively. As seen in these tables the ASHRAE standard design heat loss (HID) overpredicts the calculated design heat loss (HLC) by up to 16% for an infiltration rate of 4 air changes per hour.5.5 to 4. As the infiltration rate increases. If this were compared to a forced air system with an air temperature gradient. It is interesting to note that the percent difference in design heat losses remains about the same (-4%) for all of these cases.By increasing the panel convection coefficient by 100% (Table 16).4) so that there might be a difference of approximately 17%.3 . Some of the other parameters change slightly with significant changes in panel convection coefficient. effective radiant flux and AUST increase significantly. the floor temperature. 20 and 21 for 130. mean radiant temperature. 5. the area required for heating the space is reduced by 4. operative temperature.5% for a 150 F panel temperature. Therefore.8%. These changes need to be considered in the design process for radiant panel systems. o ••-I •n a ex. 60 4 6 Convection multiplier FIGURE 20. EFFECTS ON PERCENT RADIATION DELIVERED BY THE PANEL AS THE CONVECTION MULTIPLIER IS CHANGED . I 00 o i QI 80 - 70 - o c +J QI U L a a.100 C 0 90 - ••I +> . BTU/HR EFFECTIVE RADIANT FIELD.6 76.4 -8.SQ FT 67.2 93.8 53552.0 -21.1 29386.1 94.0 61. DIMENSIONLESS 0.6 83.7 394. F 74.7 24167.2 30481.0125 INFILTRATION AIR.4 31060.1 -16.9 94.3 46304. BTU/HR.7 76.3 -11.9 66.7 39056.8788 0.0 42508.7 28501.2 55.TABLE 19.U.4 66.4 21.6 799.0400 1.8 HEAT OUTPUT PER UNIT PANEL AREA.6 ASHRAE DESIGN HEAT LOSS.1 80.9 43.7 68.6 32575.3 -18. DEG.0 ACTUAL DESIGN HEAT LOSS.0 -12.8 -6.4 63504.9 -15.9 26364. DEG.4 88.0144 0.7 75.1 81.3 46706.50 0.8 64.3 -11.8 47.5 ROOM AIR TEMPERATURE.8 53552. BTU/HR 23696.3 58.0136 0.4 429. BTU/HR.7 PERCENTAGE RADIATION 95.8 PERCENT CEILING COVERED BY PANELS 39.9385 0.7 28174.7 28174.3 81.6 16.0638 1. F 72.2 -3.1 79.4 36417.75 1.6 63.0 687.5 -4. BTU/HR PERCENTAGE DIFFERENCE 2 CONDUCTION DESIGN HEAT LOSS 2.6 25636.7 -10. F 76.1 25815.3 27497.4 9.2 37260. EFFECTS DUE TO CHANGING AIR INFILTRATION RATES FOR A PANEL TEMPERATURE OF 130°F 0.9 34419.4 95.3 62.F 1. BTU/HR 26762.1 43453.8 74.4 72.6 69.8 40625.2 30481.8 73. BTU/HR.2 -3.00 3.7 72.8340 PARAMETER 3. -9.9 75.S.8 65. F 66.5 -4.1 -11.8 32011.1 -4.7 55.2 FLOOR TEMPERATURE.0167 0.0128 0.5 PERCENTAGE DIFFERENCE 4.8 -14. DEG.4 90.4 25.4 49742.6 -19. DEG.3 46706. AC/H PANEL AREA REQUIRED .4 73.1 -6.1 • 34889.2 62.50 2.0 7.6 94.0 10.SQ FT.4 -8. SQ FT PERCENTAGE DIFFERENCE 1 CONDUCTION DESIGN HEAT LOSS 1.0 -23.T.0133 0.7 70.1 -11.1 74.9751 0.8 t .5 -12.9 -15.00 1.7 25815.4 PARAMETER 1. SQ FT A. DEG. F -4.6 -14.2 95.00 358.0142 0.9 78.9 73.9 78.7 39056.5 75.8 63. BTU/HR PERCENTAGE DIFFERENCE 3 ACTUAL HEAT INPUT.0139 0.8 71.8 53006.4 -27.5 MEAN RADIANT TEMPERATURE.1 69.3 65.4 64.6 13.00 4.6 498.1 564.5 OPERATIVE TEMPERATURE.1 34889.7 87. 0 ACTUAL DESIGN HEAT LOSS.4 90.1 46297.0 42508.6 PARAMETER 1.3 63.4 356. EFFECTS DUE TO CHANGING AIR INFILTRATION RATES FOR A PANEL TEMPERATURE OF 150°F INFILTRATION AIR. AC/H 0.3 58.8 25644.0110 0.F 1.3 39.2 34894.4 25.3 81.8 -6.3 21.9 75.75 1.5 69.2 92.0 61.3 28172.4 308.0 40640.00 4.1 -11.0109 0.0 53550.2 95.3 95.4 95.0097 PANEL AREA REQUIRED . BTU/HR 23697.6 -9.1 -16.7 33233.9875 0.0 31072.6 ROOM AIR TEMPERATURE.9 10. F .5 70. F 66. BTU/HR EFFECTIVE RADIANT FIELD.S.1 -4.5 75.9 73.0 -8.9464 PARAMETER 3. SQ FT.1 567.4 29386.0 -23.6 87.2 74.0 94.3 -8.2 PERCENT CEILING COVERED BY PANELS 28.8 -6.8 64.5 -12.00 3.5 -4. DEG.6 -16.0 -3.3 90.4 73. DEG.8 94.9 34435.3 -8.5 94.7 82.SQ FT 95.4 26865.0100 0.5 73.1 -4.8 89.4 72.5 402.1 25813.4 8.5 83.9 63.0108 0. BTU/HR.0 94.0 -3. BTU/HR.5 13.9 -15.1236 1.2 44388.8 -12.0398 0.5 71.1 -11.3 28172.8 65.4 63504. DEG.7 -14.2 37173.8 53006.7 25813.5 -4.0105 0.0 PERCENTAGE RADIATION 95.50 0. BTU/HR PERCENTAGE DIFFERENCE 2 CONDUCTION DESIGN HEAT LOSS 2.7 77. DEG.5 16.8 32011.3 30483.0103 0.7 68.7 72. BTU/HR 26762.9 46715.7 ASHRAE DESIGN HEAT LOSS.00 1.5 OPERATIVE TEMPERATURE. F 72.3 -27.1 78.50 2.U.6 44.4 80.7 39064.3 -20.2 284.0 69. SQ FT A.2 -10.1 79.1 HEAT OUTPUT PER UNIT PANEL AREA.2 55.T.00 259.TABLE 20.4 -11. BTU/HR PERCENTAGE DIFFERENCE 3 ACTUAL HEAT INPUT.2 93..3 30483.2 37260.1425 1.7 54.9 46715.5 FLOOR TEMPERATURE.9 78.7 24621.2 74. SQ FT PERCENTAGE DIFFERENCE 1 CONDUCTION DESIGN HEAT LOSS 1.1 7. DEG.8 31. BTU/HR.5 94.7 489.4 50826.9 -15. F 74.5 MEAN RADIANT TEMPERATURE.0708 1.1053 1.4 PERCENTAGE DIFFERENCE 4 -8.0 53550.6 -19.6 29058.2 34894. F 76.9 27505.8 74.6 34.7 39064. DIMENSIONLESS 0. 6 •34899.6 25813.0 -23.9 63. BTU/HR.9 69.3 95.3 -27.2 55.5 429.7 28173.U. BTU/HR.5 46319.1 -11.8 53006.1 -4.0086 0.5 -12.0 ACTUAL DESIGN HEAT LOSS.8 74.5 46729.2 37260.1572 1. F 74.2 40658. BTU/HR HEAT OUTPUT PER UNIT PANEL AREA.6 123.2 -18.1 39072.4 73.7 123. SQ FT PERCENTAGE DIFFERENCE 1 CONDUCTION DESIGN HEAT LOSS 1. F -4t8 .5 16.1 79.9 82.4 71. F 76.2 34. SQ FT A.0085 0.3 27507.3 30485.3 -8.3 .0 41.3 95.5 PERCENT CEILING COVERED BY PANELS 22.SQ FT.2 -9.4 95. BTU/HR.7 72.1 95.1 -6.4 72.2 78.4 73.1 -3.0088 0.7 125.8 272.8 PERCENTAGE DIFFERENCE 4 -7.50 0.8 -15.5 25647.0850 1.2031 1. BTU/HR 26762.5 83.3 58.50 2.2189 1. AC/H PANEL AREA REQUIRED .9 73.4 8.9 78.0 75.7 68.3 37632.00 198.5 -4. DEG.2 47.0090 0. .5 -4.4 29386.5 MEAN RADIANT TEMPERATURE.0 305. DEG.TABLE 21.2 75.7 94.0 235.1868 1.8 32011.6 -19. BTU/HR 23696.0 42508.5 80.SQ FT EFFECTIVE RADIANT FIELD.0 29395.0 94.75 1.00 1.1 61.5 13.T.4 25.7 77.2 124.1312 1.3 -8. BTU/HR PERCENTAGE DIFFERENCE 2 CONDUCTION DESIGN HEAT LOSS 2.5 -15.1 39072.0089 0.0478 PARAMETER 3.8 -11.6 34899. DIMENSIONLESS 0.8 31079.3 21.9 ROOM AIR TEMPERATURE.9 10.2 27175.0080 INFILTRATION AIR.90. DEG.1 26. F 72. DEG.8 -6.7 -11.S. BTU/HR PERCENTAGE DIFFERENCE 3 ACTUAL HEAT INPUT.5 46729.0082 0.F 1.3 81.0 74. DEG.9 PERCENTAGE RADIATION 95.8 -15.1 -3.0 ASHRAE DESIGN HEAT LOSS.1 -16.3 PARAMETER 1.1 25813.7 125.00 4.6 87.4 70.9 370.5 OPERATIVE TEMPERATURE.0 121. F 66.5 68.1 44961.8 65. EFFECTS DUE TO CHANGING AIR INFILTRATION RATES FOR A PANEL TEMPERATURE AT 170°F 0.8 53570.0 217.5 -8.0 33629.8 53570.3 30485.2 51501.1 FLOOR TEMPERATURE.0 24.7 -14.7 28173.2 30.5 75.6 24902.4 120.0 -7.0 34445.0 7.1 -11.4 63504.8 64.00 3. 0} 01 U c. Cach) EFFECT OF A I R I N F I L T R A T I O N I N PANEL HEATING ON PERCENT DIFFERENCE I N DESIGN LOAD CALCULATIONS . 0) L at I 00 I 07 a +» c ai u -12 - -IS - L <J Q. -18 infiltration FIGURE 2 1 . 7 72. F 72.8 73.9 466.T .2 -3. SQ FT PERCENTAGE DIFFERENCE 1 CONDUCTION DESIGN HEAT LOSS 1.5 9.1 -3. DEG.5 ASHRAE DESIGN HEAT LOSS.1 34945.9 45.4 31428.4 21110. DEG.5 -4.3 80.1 -4. a l l glass-second w a l l .4 68.7 51.2 359. SQ FT A.0121 0.8 66.2 ACTUAL DESIGN HEAT LOSS. F 76. BTU/HR EFFECTIVE RADIANT FIELD.6 40759.8 76.8 -13.0 6.0 10.3 95.PANEL TEMPERATURE = 140°F 1 2 3 4 5 249.7 33.0803 1.4 PERCENTAGE RADIATION 95.3 66.3 34757.5 67.8 79.3 -7.4 28839.T. h a l f glass One w a l l .S.SQ FT 80. F Case Case Case Case Case 3. -3.3 80.6 -12.2 95.0124 0.1 39796.8 75.0 36093.6 -14.3 72.3 -8.7 65.4 72.2 64. BTU/HR PERCENTAGE DIFFERENCE 3 ACTUAL HEAT INPUT.8 24263.5 -4.2 PERCENT CEILING COVERED BY PANELS 27.1 34945.4 95.3 37737.9 OPERATIVE TEMPERATURE.1 30469.8 26762.U. F 67. BTU/HR PERCENTAGE DIFFERENCE 2 CONDUCTION DESIGN HEAT LOSS 2. F 73.0126 0.7 -9.3 33107.9 FLOOR TEMPERATURE.1 31189.7 PERCENTAGE DIFFERENCE 4 -9.1 30469. BTU/HR 19744.1 27407.0761 1.F 1.9 23671.6 80.2 -3.8 -10. h a l f glass Two w a l l s .0 PARAMETER 1.6 77.1 39796.4 ROOM AIR TEMPERATURE. BTU/HR. BTU/HR.0687 PARAMETER 0.0 11.2 -2.8 74.8 66.2 77.2 -8.wm r x -3 r—5 TABLE 2 2 .8 65. a l l glass .0937 1.3 410.8 HEAT OUTPUT PER UNIT PANEL AREA.2 -2.6 39.SQ FT. DEG.2 25650.3 69. '• -•* r^ EFFECTS OF CHANGES IN GLASS DISTRIBUTION . DEG. BTU/HR.0119 0.0 78.1 78.6 -11.4 19954.0117 GLASS DISTRIBUTION CASE NUMBER PANEL AREA REQUIRED .1013 1.2 25650.DIMENSIONLESS number number number number number T: 2: 3: 4: 5: No glass In any wall One w a l l .4 7.1 -4.. DEG.0 302. BTU/HR 22096.7 21110.4 MEAN RADIANT TEMPERATURE. a l l glass One w a l l .2 95.1 80. Also. This in turn causes the room air temperature for comfort to be reduced from 67 to 64 F.1 to 0. as the quantity of glass increases the panel area increases. as can be seen in Table 22. 150 and 170 F. the required panel area is plotted for each case shown in Table 22. Floor and Ceiling U-Factors.07 to 0. These results are given in Table 23. In each case. The radiant base case is shown as Case 2 in Table 22 and a room with no glass is given as Case 1. The wall U-factors were changed from 0.58 to 1.of glass have been considered in the radiant base case which was previously described. plotted. Case 4 is one wall with all glass and half of another wall with glass. the floor temperature also rises. The U-Factors in the radiant base case were changed and various calculations were made to determine this effect on the design heat loss. the difference between HID and HLC has been reduced by about one-half so that the ASHRAE standard design -86- . the difference between HID and HLC becomes smaller as the quantity of glass in the room increases with only a -2% difference showing up in Case 5.0 Btu/hr ft2 F. in Figure 23 the floor temperature is It is interesting to note that an 87% increase in panel area in the room results in a only a 3. and Case 5 is the room with two walls all glass. The U-factor for the glass was changed from 0.5. The variation is very much as expected in that with increased U-factors there is an increased heat loss and greater panel area required. Case 3 is one wall which is all glass.6 F increase in floor temperature. All of these were changed at one time so that an initial and a new case were compared at three panel temperatures of 130.2 Btu/hr ft 2 F and the floor and ceiling values from 0.4 . Since the panel area increases in order to make up for increased heat losses as the quantity of glass is increased.1 Btu/hr ft2 F. 5. The results from these calculations are given in Table 22 for a panel temperature of 140 F. This results in an increased floor temperature and MRT and a decreased air temperature and AUST.Changes in Wall. Likewise. As anticipated. In Figure 22. i-'a ey a en I oo I UJ z 4 CASE NUMBER FIGURE 22. PANEL AREA REQUIRED AS A FUNCTION OF THE QUANTITY OF GLASS .mî r* K e^^ ss-s-. cc o o 73.0 -.0 70.0 -r 79. FLOOR TEMPERATURE AS A FUNCTION OF THE QUANTITY OF GLASS .0 I 00 oo I UJ Q. 2 74.0 78.0 71.0 O 77.0 -• 1 2 T" T" 3 4 CASE NUMBER FIGURE 23.80.0 - 111 Q a: 75.0 - UJ 76.0 72. 4 41536.6 -14.1 35622.5 10.0105 0.1176 1.3 68.8 7.7 64.5 PERCENTAGE RADIATION 95.0523 1.4 40542.4 -4.9 72.6 40403.9 68.9 94.0109 0.2 FLOOR 0.3 197.0 33. BTU/HR 26762.4 125.2111 1.3 ASHRAE DESIGN HEAT LOSS.2 WALL 3 0.6 40403.3 95.• . DEG.1 -2.4 77.6 24459.0 630.07 0.0 23678.8 25647.7 64. DEG.4 95.7 66.3 24735.SQ FT 66.7 397.2 " -2.1338 1.1 22. . DEG.2 WALL 4 0.10 .0 125.2 -2.7 -4. D E C F PANEL AREA REQUIRED .0086 PANEL TEMPERATURE.7 MEAN RADIANT TEMPERATURE.5 -14.3 25654.0 590.7 67.8 68.5 10.2 95. F SURFACE INITIAL U NEW U WALL 1 0.7 PARAMETER 1.4 40542.0 630.5 -14.7 ROOM AIR TEMPERATURE.9 72.4 72. DIMENSIONLESS 0.2 40374.9 305.0 79.2 -2.3 95.0 610.4 -11.1 0. F 74.2 38435.34 0.1 -2.6 95.9 HEAT OUTPUT PER UNIT PANEL AREA.3 -10.7 44.0 60.0089 0. DEG.9 FLOOR TEMPERATURE.0 79.3 -11.8 23674.4 -4.1 0.0136 0.7 -8.0142 0.7 7.4 72.* §'••'• •>•:••: s t< '.10 CEILING 0.3 PERCENT CEILING COVERED BY PANELS 40.60 WALL 2 0.SQ FT. BTU/HR PERCENTAGE DIFFERENCE 2 CONDUCTION DESIGN HEAT LOSS 2.5 10.9 25645.8 25654..8 -7.0 610. S Q F T A.6 37890.7 -4.3 258.7 28.F 1.4 65.8 24005.7 -4. EFFECTS OF INCREASED U-FACTORS ON RADIANT HEATING PANEL PERFORMANCE U-FACTOR CASE OLD NEW OLD NEW OLD NEW 590.4 72.7 75.0 79.7 66.» . F 66.2 95.6 74.2 -11.4 74. BTU/HR.07 0.T. BTU/HR PERCENTAGE DIFFERENCE 3 i 00 I ACTUAL HEAT INPUT.1 0.0390 1. BTU/HR PERCENTAGE DIFFERENCE 4 EFFECTIVE RADIANT FIELD.6 66.S.8 76. BTU/HR. F 77.5 65.4 41536.6 -8.8 25647. DEG.5 77. BTU/HR 23667.7 -4.1947 PARAMETER 3.0 359. F 72.U.8 26762.8 26762.6 -7.9 25645.8 ACTUAL DESIGN HEAT LOSS.7 546.0 35611.0 -10.4 41536.8 37107.V! TABLE 23.2 -2.5 OPERATIVE TEMPERATURE. SQ FT PERCENTAGE DIFFERENCE 1 CONDUCTION DESIGN HEAT LOSS 1.7 64.9 76.9 7.7 35605. BTU/HR.2 40374. BTU/HR PERCENTAGE DIFFERENCE 3 ACTUAL HEAT INPUT.0937 1.6 74. BTU/HR PERCENTAGE DIFFERENCE 2 CONDUCTION DESIGN HEAT LOSS 2.SQ 82.0 23579.0 26726.0124 0.0 24263.3 ROOM AIR TEMPERATURE.1191 1. DEG.2 -5.2 ASHRAE DESIGN HEAT LOSS.4 17658.3 80.1217 PARAMETER 3.4 25826.8 -3.7 -4. BTU/HR.4 66.6 FLOOR TEMPERATURE.0 67. FT*FT 20*20 30*30 40*40 40*20 30*15 PANEL AREA REQUIRED .9 300.9 197.8 PERCENTAGE RADIATION 94.5 72.7 74.6 29.1047 1.0822 1. F 77. BTU/HR 12901.6 PERCENTAGE DIFFERENCE 4 -8.9 95. DI MENS I ONLESS 0.6 25650.6 43.1 81. DEG. F 66.0 -11. BTU/HR.8 7.4 42048.F 1.0 17178.6 -12.8 -11.5 72. SQ FT 162.0126 0.4 72.TABLE 24.9 68.3 80.7 33.4 37571.6 7.8 -3. DEG.5 8.7 66.1 39711. DEG.8 77.6 -3.8 74.1 16284. BTU/HR EFFECTIVE RADIANT FIELD.5 66.5 8.2 OPERATIVE TEMPERATURE.0124 0. SQ FT A.0 68.6 72.9 67.0 66.5 MEAN RADIANT TEMPERATURE.0 17178.3 -10.8 14121.U.6 95.4 -11.0122 0.8 PARAMETER 1.2 26762.7 -4. F 72. F .4 -2.2 77.2 468.0 77. EFFECTS OF CHANGES IN ROOM LENGTH AND WIDTH WITH A 140°F PANEL TEMPERATURE ROOM LENGTH*WIDTH.2 -5.1 7.5 -7.S. BTU/HR 14659.3 82. DEG.6 24454.0 ACTUAL DESIGN HEAT LOSS.0126 PERCENTAGE DIFFERENCE 1 CONDUCTION DESIGN HEAT LOSS 1.4 -2.9 PERCENT CEILING COVERED BY PANELS 40.7 15571.9 23671.3 37.7 -9.6 -8. F 73.3 95..6 ' -3.2 94.1 39711.4 77.5 68.1 -12.4 25826.0 74.6 302.7 13384.7 14121.1 36820.8 HEAT OUTPUT PER UNIT PANEL AREA.6 25650. BTU/HR.SQ FT.T. The radiant base case room was modified to have a ceiling height between eight and twenty five feet.. In Table 24. x 40 ft. more of the walls intercept the radiant energy thus increasing the AUST. This tendency is not great (3. 5. 5.. x 20 ft.5 . In Table 25. as the height is increased. but it is an important trend. 5.3% at 15 F) .000 sq. x 30 ft.5.7 . results for four square buildings are tabulated.procedure more closely predicts the required heat input. building ) the radiant heating system.6% for 40 x 40). The results from these calculations are presented in Table 26. x 20 ft. There are two important trends to observe from these results. ft. There is a slight tendency for the percent difference between HLD and HLC to increase with milder climates (-3. This in turn causes the second trend to occur in that the difference between HLD and HLC decreases because of more heat conduction through the walls.. This does not -91- . all 9 ft. In order to see what effect outside design temperature had on the design load calculation.Effect of Changes in Room Length and Width. 30 ft.6 . five other outside design temperatures were used and these results are given in Table 27.7% for 20 x 20 and 5. are the results for five different size rooms: 20 ft.Changes in Room Height.5. high. more panel area is required to counteract the increased room heat loss and because of changing room geometry. Also illustrated here is the fact that a square room or building will tend to be over sized if the ASHRAE standard design load is used. 40 ft.6% at -5 F to -5. This decrease in the difference between HLD and HLC as room height is increased is illustrated by the plot shown in Figure 24. x 15 ft. Again. 40 ft. as the building becomes larger the ASHRAE standard design procedure (HLD) tends to oversize (6% for a 10.Changes in Outside Design Temperature. First.5.. and 30 ft. The important thing to notice about these results is that as the room size increases the ASHRAE standard design load procedure tends to increasingly over predict the required heater size. U.6 25650.7 66.3 80.6 302.0122 0.6 OPERATIVE TEMPERATURE.0126 0.7 -9.F 1.4 77.8 -12. F 73.6 -5.0566 PARAMETER 3.7 ROOM AIR TEMPERATURE. F .9 23671. DIMENSIONLESS 0.8 ASHRAE DESIGN HEAT LOSS.2 26762.5 68.2 468.3 95. SQ FT 162.6 14121.6 96.0 39711.7 74. DEG.8 FLOOR TEMPERATURE.4 72.0 176195.7 33.0 200952.0 66.4 188495.0 77. EFFECTS OF CHANGING ROOM SIZE WITH A 140°F PANEL TEMPERATURE ROOM LENGTH*WIDTH.1 -3.4 188495. DEG.4 42048.3 22.0 ACTUAL DESIGN HEAT LOSS.8 77. F 66.0 24263.0 -11.4 66.3 80.3 PERCENTAGE RADIATION 94. DEG.1 77. F 72. SQ FT A. BTU/HR PERCENTAGE DIFFERENCE 3 ACTUAL HEAT INPUT.6 29. BTU/HR 14659.4 -12.6 72.0 PARAMETER 1.S.3 -10.6 177908. F 77.6 " -6.5 72.0937 1.0124 0. BTU/HR 14121. BTU/HR.2 13384.8 74.6 -6.TABLE 25.2 MEAN RADIANT TEMPERATURE.2 PERCENTAGE DIFFERENCE 1 CONDUCTION DESIGN HEAT LOSS 1.6 8.T.1 PERCENTAGE DIFFERENCE 4 -8.3 67.SQ FT.1 7. BTU/HR 82.0822 1.1 -3.0120 EFFECTIVE RADIANT FIELD. BTU/HR 12901.9 2285.5 8.1191 1.7 -4. BTU/HR PERCENTAGE DIFFERENCE 2 CONDUCTION DESIGN HEAT LOSS 2. DEG.7 -4.9 HEAT OUTPUT PER UNIT PANEL AREA. DEG.1 -12.6 25650.2 39711.0 73.0 69.9 68.0 37571.6 -5.9 95.4 PERCENT CEILING COVERED BY PANELS 40. BTU/HR.5 8.6 -11. FT*FT 20*20 30*30 40*40 100*100 PANEL AREA REQUIRED .4 36820. 2 77.4 74.0 23528.0 58212.SQ FT 81.0983 1.5 -11. SQ FT PERCENTAGE DIFFERENCE 1 CONDUCTION DESIGN HEAT LOSS 1.0 274.5 47707.TABLE 26.6 27642.5 8. DEC.3 95. DEG.7 68.4 22280. BTU/HR.3 PERCENT CEILING COVERED BY PANELS 30.4 72.6 80.7 31529.5 7.3 95. DEG.7 68.0 32659. DEG.3 -1. BTU/HR 24796.0124 0.0 33584.0123 0.5 74.1 81. BTU/HR PERCENTAGE DIFFERENCE 4 EFFECTIVE RADIANT FIELD.6 -8.5 -3.3 HEAT OUTPUT PER UNIT PANEL AREA.7 72.7 23684.4 -0.0 15.3 95.SQ FT.8 -12.0124 0. F .8 75.2 38556.5 33.6 78.4 28728.1 67.2 65.7 74.0 51235.0127 PANEL AREA REQUIRED .2 -12.4 MEAN RADIANT TEMPERATURE.7 29847.9 -3.9 -3.5 -5.F 1.2 80. BTU/HR 21765.0 40.2 66.5 47707.6 72.3 95.4 OPERATIVE TEMPERATURE.2 66.0 12. EFFECTS OF CHANGING ROOM HEIGHT WITH A ROOM HEIGHT.8 -3.2 -11.0994 1.0124 0.9 8.5 -3.S.5 36. -3.7 31529.0 48384.8 -10.6 77.8 81.2 -12.7 35307.6 72.6 28675.8 .5 68.7 -5. BTU/HR PERCENTAGE DIFFERENCE 3 ACTUAL HEAT INPUT.0983 1.6 37271.0989 1.0971 PARAMETER 3.7 66.2 77.3 8.6 68.6 367.8 74.2 80.2 301.3 45149.3 95. DIMENSIONLESS 0.U.0 9.4 -0.1 -3. F 72. FT °F PANEL TEMPERATURE 8.0 556.4 23528.8 ASHRAE DESIGN HEAT LOSS. BTU/HR PERCENTAGE DIFFERENCE 2 CONDUCTION DESIGN HEAT LOSS 2.1 677.0990 1. BTU/HR.5 -5.5 66.8 26762.8 70.8 432.1 70. SQ FT A.0124 0.5 -12.1 -3.2 25728.0 -8.0126 0. F 73.0 61.8 54824.4 -6.5 42498.2 57951. BTU/HR.3 -1.0 10.1 -9.2 57951.6 27642.7 66.5 72. F 77.1 8.T.3 9.6 ROOM AIR TEMPERATURE.0 77.8 PERCENTAGE RADIATION 95.0995 1. F 66.5 72.9 FLOOR TEMPERATURE.3 7.0 20.6 74.2 25728.3 95.6 77.9 323.9 -12.5 25332.7 81.8 PARAMETER 1.9 -8.8 24342.5 74.0 25.9 48.6 37271.0 ACTUAL DESIGN HEAT LOSS.1 26159. DEG. Conduction Design Heat Loss 48000 - +J CD W I 43000 - 0 HJ o a x 38000 33000 h 28000 23000 10 Room height FIGURE 24. 20 15 (ft) EFFECT OF ROOM HEIGHT CHANGE ON DESIGN HEAT LOSS FOR RADIANT PANELS 25 .ASHRAE Design Heat Loss 53000 L H s 0 = HLC .63000 58000 A = HLD . 5 72. BTU/HR 26437.4 23012.4 24899. F .0141 0.2 PANEL AREA REQUIRED .3 -3.8 74.2 67.4 -4.1 21675.7 FLOOR TEMPERATURE. BTU/HR I PERCENTAGE DIFFERENCE 4 -11.TABLE 27.U. DEG.4 24899.7 74.0663 PARAMETER 3.0 3. DEG. DEG.1 -4.0144 0.0 24005.9 28675.054.4 72.0142 0.8 6.3 41.6 -3. F 66.6 72.7 66.5 26762.5 MEAN RADIANT TEMPERATURE.4 68. EFFECTS DUE TO OUTSIDE DESIGN TEMPERATURE CHANGES WITH A 130°F PANEL TEMPERATURE OUTSIDE DESIGN TEMPERATURE.7 66.5 7.3 24996.5 22302.3 -12.3 -10. BTU/HR.6 40.4 PERCENT CEILING COVERED BY PANELS 44.7 66.0491 1. F 77.2 8. BTU/HR PERCENTAGE DIFFERENCE 2 CONDUCTION DESIGN HEAT LOSS 2.1 66.3 -10.1 21125.4 26019. SQ FT PERCENTAGE DIFFERENCE 1 CONDUCTION DESIGN HEAT LOSS 1.7 324.T.3 95.7 26787.8 -5.0 0. F 72.8 24706.5 25654.9 7. BTU/HR 29736.5 ASHRAE DESIGN HEAT LOSS.4 -11.3 95.8 23343. BTU/HR.0 27877. DEG.0604 1.7 22975.7 68.7 26787.F 1. F 75.4 28675.0 OPERATIVE TEMPERATURE.3 HEAT OUTPUT PER UNIT PANEL AREA.SQ FT 66.3 6.6 68.0 38.8 67. DEG.7 -4.0 76.0 15.3 -10.8 76.8 -3.3 -4.0145 EFFECTIVE RADIANT FIELD.6 26634.0 24160.5 359.1 -11.3 77.3 95.4 PERCENTAGE RADIATION 95.8 299.0437 1.4 -10.8 69.7 77.9 36.0143 0.6 -11.0 399. SQ FT A.5 66.7 349.5 7. BTU/HR.SQ FT.0142 0. DIMENSIONLESS 0.0523 1.3 95.8 19512.0 74.0 10.3 25654.1 -12.7 66.4 74. BTU/HR PERCENTAGE DIFFERENCE 3 ^ ACTUAL HEAT INPUT.3 95.8 -5.0 374.6 1.0 5.S.3 -4.3 72.5 19992. D E C F -5.3 -3.8 66.1 21125.7 66.1 -4.1 33.4 76.9 -10.2 68.0 ACTUAL DESIGN HEAT LOSS.4 72.4 23667.6 74.0 PARAMETER 1.1 ROOM AIR TEMPERATURE.4 23012.3 -10.5 68.3 21243. and is apparently due to reduced infiltration and conduction losses at the higher outside temperatures. The other variables in Tables 13 and 29 are quite similar so that this special type of application of panels does not alter the conclusions from the single -96- . This is due to the walls intercepting more of the radiant energy in the 6 panel case than in the single panel case due to changing angle factors. It is interesting to note that the floor temperature did drop by 1 F in going from one to six panels and the AUST increased by a slight (less than 1 F) amount. This is only about 4% different than the results shown in Table 13 for the single panel radiant base case. These cases are for a 15' x 15' x 8' room with three inside walls and one outside wall with half glass. The radiant base case was used again except that 2. This was done to determine what effect panel distribution has on the design heating load. There is no apparent difference between HID and HLC as far as the design loads are concerned.8 . 5. The results from these calculations are given in Table 28.Changes in Number of Panels.4. Several cases were run with a perimeter (narrow panel running parallel to the outside wall) radiant panel system and these results are given in Table 29. The U-factor for the floor and ceiling were the same as the radiant base case. This is apparently due to the proximity of the radiant heating surface to the cold surface or wall resulting in higher convective losses. All of the calculations presented to this point have been for a single panel located in the center of the space or room at ceiling height. The panel area required and difference in design heating loads did not change significantly.appear to be a significant trend. The radiant panel was 36"wide and ran parallel to the outside wall with the window. This is not a significant trend considering all of the unknown variables that can enter into consideration in the actual case.5. and 6 equal area panels were used for supplying the radiant heat to the room. 5.5.9 .Perimeter Panel System. DIMENSIONLESS 0.S.1068 1.5 33.4 25650.1 FLOOR TEMPERATURE.3 -9.3 95.1 -4.2 -3.0124 0.SQ FT.5 68.4 26762. SQ FT PERCENTAGE DIFFERENCE 1 CONDUCTION DESIGN HEAT LOSS 1.4 -3.F 1. F . F 66. BtU/HR.0 77.-"-v>a f- CHANGES AS A RESULT OF THE NUMBER OF RADIANT HE NUMBER OF PANELS NG PANELS FOR A 1 4 0 ° F PANEL TEMPERATURE 1 2 4 6 302.0954 1. BTU/HR .0 PERCENTAGE DIFFERENCE 4 -9.1 25649.0937 1.4 72.6 25850.6 7.4 PERCENT CEILING COVERED BY PANELS 33. BTU/HR PERCENTAGE DIFFERENCE 2 CONDUCTION DESIGN HEAT LOSS 2. DEG.4 24263.2 301. BTU/HR.3 -11.5 25650.0 OPERATIVE TEMPERATURE. DEG. BTU/HR PERCENTAGE DIFFERENCE 3 ACTUAL HEAT INPUT.5 7.0126 0. F 72.4 26762.2 -3.5 -11.3 95.4 7.8 PARAMETER 1.0 77.6 24450.1 25649.5 7.8 ASHRAE DESIGN HEAT LOSS.6 -11.8 68.5 80.1 -4.0124 0. F 74.SQ FT EFFECTIVE RADIANT FIELD.3 -8.4 ROOM AIR TEMPERATURE. DEG.6 73. DEG. HEAT OUTPUT PER UNIT PANEL AREA.4 81.5 68.4 24263.3 80.2 -4.7 66.5 33.7 301.4 -3.0126 PANEL AREA REQUIRED .0 81. ' ^ p.4 24460.6 PERCENTAGE RADIATION 95.7 MEAN RADIANT TEMPERATURE.7 68. F 77.6 -11.2 -4.4 26762.0 25840.7 66.6 -8. BTU/HR 26762.8 23682.0 77.1 23665.6 25850.1075 PARAMETER 3.4 95.6 33. SQ FT A. BTU/HR 23671.4 72. BTU/HR.T.7 66.4 ACTUAL DESIGN HEAT LOSS.„sg gp:"^ p . DEG.8 74.6 23681.2 73.0 25840.U.8 301.4 72. 2808 t.4 OPERATIVE TEMPERATURE.0 185. DEG.5 -9. SQ FT A.5 69.1 6390.3 -0.0 6171.3 -0.3 71.8 -9. DEG. BTU/HR PERCENTAGE DIFFERENCE 3 ACTUAL HEAT INPUT. BTU/HR.9 137.9 67.4 FLOOR TEMPERATURE. F .8 94.9 ACTUAL DESIGN HEAT LOSS.5 40.U.5 5.7 ' 154.2993 1. BTU/HR.F 1. F 71.S. SQ FT PERCENTAGE DIFFERENCE 1 CONDUCTION DESIGN HEAT LOSS 1.2 -2. D E C F PANEL AREA REQUIRED .5 -3.5 PARAMETER 1. RESULTS FOR A PERIMETER RADIANT PANEL HEATING SYSTEM .4 5.3176 PARAMETER 3.1 5.9 67.2 ASHRAE DESIGN HEAT LOSS.4 75.SQ FT.8 -9.1 72. DEG. F 72.36" WIDE PANEL .9 PERCENTAGE RADIATION 94.0085 0.9 17.3 -0.8 94.3 5768.5 69.3 r0. BTU/HR 6397.SQ FT EFFECTIVE RADIANT FIELD.7 6192.9 6397.5 69.1 0.0 18.2 71.1 PERCENTAGE DIFFERENCE 4 -3.4 75.9 6397.3 6210. DIMENSIONLESS 0.FLOOR AND CEILING AT RADIANT BASE CASE CONDITIONS 175. BTU/HR.9 MEAN RADIANT TEMPERATURE. F 67.0 45.0 6381.8 PERCENT CEILING COVERED BY PANELS 20. BTU/HR 5770. DEG.0 6397.TABLE 29. BTU/HR HEAT OUTPUT PER UNIT PANEL AREA.8 6381.1 0.1 72.0 6397.0 180.2 145.2 5769. BTU/HR PERCENTAGE DIFFERENCE 2 CONDUCTION DESIGN HEAT LOSS 2. F 75.0081 PERIMETER HEATER TEMPERATURE.0089 0. DEG.T.4 ROOM AIR TEMPERATURE.0 42.1 6390.15' x 15' x 8' ROOM WITH THREE INSIDE WALLS AND ONE OUTSIDE WALL WITH HALF GLASS . 5. two different room heights are compared for forced air and radiant heating systems. Figure 25 shows radiant ceiling panels at 120 F (rad-120) and at 180 F (rad-180) compared with forced air heating systems with an air temperature gradient of 0.5 F/ft (con-1.50). Keep in mind that these results are fo'r a nine feet high room and one-half air change per hour. In Figure 26. The con-25 and rad-25 are for the same variables except that the ceiling is 25 feet high.6 Comparison of Forced Air and Radiant Ceiling Panels. The con-8 case is for an 8 feet high ceiling forced air system and rad-8 is for an 8 feet high ceiling radiant system. The base case which was used for this was the same as that previously described except that it has a room height of 8 feet and the outside design temperature was selected to be 10 F. From this figure it appears that the increased infiltration heat loss. The room height of 8 feet was chosen for this case since the radiant floor type system is commonly applied to residential structures. This is due to more of the radiant energy falling on the walls as the height of the building is increased. The radiant heated floor type of system has also been simulated. Several of the situations from forced air and radiant ceiling panel systems were compared and these are shown in Figures 25 and 26. These results show the same trends as discussed above except that panel heating system design loads become equivalent to the forced air design loads and the ASHRAE standard design heating loads as long as room air temperature gradients are considered.7 Heated Floor Cases.75) and a gradient of 1. 5. The 10 F outside temperature was selected since the floor temperature is lini- . including an air temperature gradient in the forced air cases.75 F/ft (con-0. is not enough to overcome the effect of an increased AUST in the radiant case (Bar-Conduction 2 in Figure 25).panel calculations. 50) CONDUCTION 2 FIGURE 25. COMPARISON OF THE FORCED AIR AND RADIANT SYSTEMS AT SELECTED SETS OF CONDITIONS .C=CONVECTIVE .75) &?Z\ C(1. R=RADIANT OH I \ fc O O o o o I en V) 3 X R(120 1771 ASHRAE R(180 F) fVÎ CONDUCTION 1 C(0. -. R=RADIANT 80.0 £* 60.0 4 X \ m o o 2 o I (A 50. COMPARISON OF THE FORCED AIR AND RADIANT SYSTEMS FOR TWO ROOM HEIGHTS pv i^ .-v.0 70.q C=CONVECTIVE .0 - I 20.0 4 Q 30.0 10.a F* v.0 0.0 C(8 f t ) 1771 ASHRAE R(8 f t ) IVÎ CONDUCTION 1 C(25 f t ) R(25 f t ) U77X CONDUCTION 2 FIGURE 26.0 40.fK^Vi «s*. 5 to 1. Next. The percent change in design load only increased an insignificant amount (1/2%). the outside design temperature was varied between 5 F and 20 F to indicate its effect on the design heat loss and these results for an 84 F floor temperature are given in Table 31. This results also in the percent radiation from the heated floor -102- . From the data in Table 30 it can be seen that the percent difference between HID and HLC is constant at about -7%. this is only 1. In Table 32. These results are presented in Table 30.25 air changes per hour for the base configuration room with an 84 F floor temperature. For this case. The trend from these calculations is that as the climate becomes milder HLD and HLC begin to diverge. the infiltration rate was varied from 0. At 1. The room air temperature remains constant at about 70 F and the MRT was approximately 73 F.9) in the deviation from the ASHRAE standard design procedure.25 ACH the floor is 100% active with heating surface.many floor temperature variations were available. this heat loss would be taken into account. There are reductions in room temperature and an increase in MRT and a resulting increase in the AUST.07 and 0. However. In Table 33.5% for an outside temperature change from 5 F to 20 F. the U-factor for the floor was changed between 0. the floor temperature was varied between 81 F and 85 F and the required floor area for heating was calculated assuming a uniform and constant floor temperature.9 to -7.ited to 85 F for comfort requirements (See Annotated Bibliograpy-Appendix B) and with a 3 F outside temperature not.15 Btu/hr ft^ F while the floor temperature was maintained at 84 F. In the desipn process. The actual heat input is lower than HLD because most of the floor is covered with radiant heating surface and no loss from the floor to the surroundings is considered for the heated area. It is seen that there is a slight increase (-6. The other variables in the calculation (except actual heat input and floor area) are affected very little by this change. 4 681.9 -6.F 1.0 22386.9 69.3 2.0 -8.5 20835. BTU/HR. BTU/HR.8 20842.1 73.2 PERCENT FLOOR 97.0 70.6 71.8 17561.7 26.0 17326.8 66.5 -6.6 71.1 -8.1 61.3 24.9 18087. BTU/HR 22386.0 22386. BTU/HR 20536.7 815. SQ FT PERCENTAGE DIFFERENCE 1 CONDUCTION DESIGN HEAT LOSS 1.8 724. DEG.0 82.8 762.9 -7.6 84.9 MEAN RADIANT TEMPERATURE.6 -21.8 20591.3 90.6 2.9 -6.7606 1.5 2. BTU/HR PERCENTAGE DIFFERENCE 3 ACTUAL HEAT INPUT.T.9 -6.1016 0.1170 0.0 66.5 20835.3 -22. F 71.9 -6. DEG. SQ FT A.2 2.0 -7.5 20842.8 80.7 HEAT OUTPUT PER UNIT PANEL AREA.5 2.7252 1.0 84.TABLE 30.9 72. F 69. RESULTS FOR HEATED FLOORS .0 ROOM AIR TEMPERATURE.9 -7. BTU/HR PERCENTAGE DIFFERENCE 4 20816.3 62.SQ FT 19. F 73.0 72.0 22386. DIMENSIONLESS 0.0 875. F 67.9 20846.0 22386.7504 1.SQ FT.1 COVERED BY PANELS EFFECTIVE RADIANT FIELD.7378 1.3 -8.7734 PARAMETER 3.6 -20. BTU/HR PERCENTAGE DIFFERENCE 2 CONDUCTION DESIGN HEAT LOSS 2.3 20830.8 20638.6 CEILING TEMPERATURE.1 17767.9 70.3 20830.6 71. DEG.0 ASHRAE DESIGN HEAT LOSS.5 75.6 71.0 70.7 61.7 66.6 66.0 ACTUAL DESIGN HEAT LOSS.8 OPERATIVE TEMPERATURE.6 -20.9 -6.5 PARAMETER 1.0 -19.7 62.0 PANEL AREA REQUIRED . DEG.0 85. DEG.5 20825.S.8 20612.BASE CASE HEATED FLOOR TEMPERATURE.8 69.1086 0.0967 0.U. F 69.9 -7.8 21.7 69. BTU/HR.9 69.5 23.9 17917.8 20564.2 PERCENTAGE RADIATION 63.2 73.5 20825.4 -6.0913 . D E C F 81.4 -6.0 83.6 69. 9 70.9 20830.6 PERCENT FLOOR 84.7606 1.5 -8. F 5.5 19115.7 66.7 69.9 67. BTU/HR 22286. F .4 24. BTU/HR 25.9 -7.6 71. F 69. F 73.0 70.5 15260. EFFECTS OF OUTSIDE AIR TEMPERATURE CHANGE FOR A HEATED FLOOR AT 84°F OUTSIDE DESIGN TEMPERATURE DEG.0 10. BTU/HR PERCENTAGE DIFFERENCE 4 COVERED BY PANELS EFFECTIVE RADIANT FIELD.F 1.0 1.9 20612.0977 0.8 724.4 -7.7 61.5 75.1 19258.8 22545.6 71.9 72.0947 0. DIMENSIONLESS 0.4 17917. BTU/HR.3 -8.5 ASHRAE DESIGN HEAT LOSS.0 22386.7 -19. DEG.9 -8.0 18942. SQ FT PERCENTAGE DIFFERENCE 1 CONDUCTION DESIGN HEAT LOSS 1. DEG.0 -19.5 2.9 16598.5 -8.6 66. DEG.5 19115.0 15.0 755.1 70.2 69.0989 PANEL AREA REQUIRED .5 -6.7523 1. F ROOM AIR TEMPERATURE.9 20830.5 24.7441 PARAMETER 3.TABLE 31.4 17273.7 24.SQ FT.1 17400.8 18943. BTU/HR PERCENTAGE DIFFERENCE 2 CONDUCTION DESIGN HEAT LOSS 2.6 -7.9 -6.2 2.1 72.S. DEG.0 ACTUAL DESIGN HEAT LOSS.5 -6.T.4 681.9 70.7 61.6 636.1 22545.1 -20.0 80.7 72. DEG.4 PERCENTAGE RADIATION 61.1 17400.0 20664.5 2.2 MEAN RADIANT TEMPERATURE.9 -6. F 71.7 61.U.1 -20. BTU/HR PERCENTAGE DIFFERENCE 3 ACTUAL HEAT INPUT. BTU/HR.0 CEILING TEMPERATURE.7 HEAT OUTPUT PER UNIT PANEL AREA. SQ FT A. BTU/HR 24108.5 71.7 70.7713 1.0967 0.5 OPERATIVE TEMPERATURE.3 66.2 PARAMETER 1.0 20.9 -7.0 70. DIMENSIQNLESS 0. DEG.5 ASHRAE DESIGN HEAT. BTU/HR PERCENTAGE DIFFERENCE 4 F EFFECTIVE RADIANT FIELD. F . F 70.9 PERCENT FLOOR COVERED BY PANELS 80.SQ FT. EFFECTS OF FLOOR U-FACTORS ON A HEATED AT 84°F 0.6 PERCENTAGE RADIATION 61. SQ FT PERCENTAGE DIFFERENCE 1 CONDUCTION DESIGN HEAT LOSS 1.9 72.0 70.9 -7.3 -7.8 20830.8 OPERATIVE TEMPERATURE. BTU/HR PERCENTAGE DIFFERENCE 2 CONDUCTION DESIGN HEAT LOSS 2.9 -7. F 71. DEG.6 66.6 71.7 61.7606 1. DEG.9 20830.9 72.5 22372.07 o.7692 PARAMETER 3.0961 0.0952 FLOOR U-FACTOR.2 -6.2 2.0 24141. BTU/HR.8 61.5 81.io 0.0 27066.9 24952. F 72. LOSS.9 17917.6 83. BTU/HR PERCENTAGE DIFFERENCE 3 ACTUAL HEAT INPUT.3 -7.0 69.0 ACTUAL DESIGN HEAT LOSS.TABLE 32. BTU/HR.4 734.7 24.U.9 MEAN RADIANT TEMPERATURE.2 2.5 -20.4 66.T.0 24928.5 2. D E C ROOM AIR TEMPERATURE.8 22239.SQ FT 24.2 PARAMETER 1.9 -7.9 -7.0 24928.7 18515. BTU/HR. DEG.8 -31.F 1.6 71.9 69. BTU/HR 20612. BTU/HR 22386.9 -7.0 -24.9 69.0967 0.1 66. BTU/HR.2 747.9 69.15 724. SQ FT A.7640 1.2 -6.F PANEL AREA REQUIRED .9 18160.8 CEILING TEMPERATURE.7 24.SQ FT.1 HEAT OUTPUT PER UNIT PANEL AREA.S.5 22372. DIMENSIONLESS 0.0947 0.6 69.8 PERCENTAGE DIFFERENCE 1 CONDUCTION DESIGN HEAT LOSS 1.9 -8.6 67.7 71.6 71.5 57.1 67.9 20830.0 -19.5 86.9 72.1 99.7 25.0 ASHRAE DESIGN HEAT LOSS. BTU/HR PERCENTAGE DIFFERENCE 3 o ON 0.6 71.3 25.9 26598.U. D E C F ROOM AIR TEMPERATURE. BTU/HR 22386.0 21337. BTU/HR.1 71.9 24734.5 PERCENT FLOOR COVERED BY PANELS 80.4 777.0 28704. SQ FT 0.4 23955. BTU/HR.SQ FT 24.4 26574.6 75.9 17917.TABLE 33.4 5.8 -6. F 71. DEG.T.9 PERCENTAGE RADIATION 61.7 -19.0 70.5 25.75 1.9 73. BTU/HR 22302. SQ FT A.0944 0.0 24492.4 26574.1 2.6959 1. BTU/HR PERCENTAGE DIFFERENCE 2 CONDUCTION DESIGN HEAT LOSS 2.9 25570. EFFECTS DUE TO INFILTRATION FOR A HEATED FLOOR AT 84°F FOR A 30' x 30" x 8' ROOM INFILTRATION RATE.7 22984.S.4 19661. DEG. BTU/HR.5 22800.SQ FT.6378 1.8 -7.0967 0. BTU/HR 20612. DEG.4 OPERATIVE TEMPERATURE.8 -6.7606 1.8 74.5 67.0946 PERCENTAGE DIFFERENCE 4 EFFECTIVE RADIANT FIELD.7 CEILING TEMPERATURE.5 22800.4 4.8 I -20.6 837.9 20830.7 68.8 -7.1 -7. F 72.0 ACTUAL DESIGN HEAT LOSS.9 MEAN RADIANT TEMPERATURE. AC/H PANEL AREA REQUIRED .0 -9.5842 PARAMETER 3.0 -7.7 59.F 1.2 68. F 70.9 -10. DEG.9 -6.0 -7.5 896.8 -19.9 69. F .9 -6.3 PARAMETER 1.5 66.6 HEAT OUTPUT PER UNIT PANEL AREA.4 93.50 ACTUAL HEAT INPUT.4 24734.2 3.00 1.5 55.25 724. 9 173.6 71.9 5433. DEG. BTU/HR. BTU/HR 5775.SQ FT.1 -19.0.7819 1.9 5775.0 68.9 82. F 71.1 -18.5 2. F -5.2 5434. BTU/HR. F 69.1045 0.8 72.T. DEG.9 59.3 27.9 5775. BTU/HR PERCENTAGE DIFFERENCE 4 EFFECTIVE RADIANT FIELD.0 84.1 2.5 71.0 197.SQ FT 23.2 4672. DEC F ROOM AIR TEMPERATURE.4 25.6 5324. DEG.0986 PERCENTAGE DIFFERENCE 2 CONDUCTION DESIGN HEAT LOSS 2.9 ' ! .9 4617.2 2.9 -7.5 PERCENT FLOOR COVERED BY PANELS 87.7 5U33.2 69. BTU/HR -5. BTU/HR 5318. F 72.1 HEAT OUTPUT PER UNIT PANEL AREA. BTU/HR.9 70.9 -5.U 5434. BTU/HR o i PERCENTAGE DIFFERENCE 3 ACTUAL HEAT INPUT.2 CEILING TEMPERATURE.9 -5.7988 PARAMETER 3. 67.2 69.1112 0. D E C F PANEL AREA REQUIRED .2 69.8 FLOOR TEMPERATURE.pw-jsaj TABLE 3 4 . P:'-'' HEATED FLOOR CASES ROOM 3 INSIDE WALLS AND 1 OUTSIDE WALL WITH HALF GLASS FOR A 1 5 ' x 1 5 ' x 8 ' 83.9 -5.2 5434.9 72. DEG.7 184.0 4720.S.1 -20.F 1.3 59.3 5329.U. DIMENSIONLESS 0.7634 1.0 68.5 ASHRAE DESIGN HEAT LOSS.0 85.7 -7.9 69. SQ FT PERCENTAGE DIFFERENCE 1 CONDUCTION DESIGN HEAT LOSS 1.9 PARAMETER 1.3 PERCENTAGE RADIATION 60.8 OPERATIVE TEMPERATURE.2 77.8 -7. SQ FT A.4 5434.8 -5.0 MEAN RADIANT TEMPERATURE.9 ACTUAL DESIGN HEAT LOSS. 4 PERCENT FLOOR COVERED BY PANELS 96.0 6091.0 -7.1 ACTUAL DESIGN HEAT LOSS.F 1.1022 0.0 67.4 6090.8152 1. DEG.SQ FT. BTU/HR. F 69. D E C ROOM AIR TEMPERATURE.0964 FLOOR TEMPERATURE.0 85.2 203.0 -7.9 5317.8385 PARAMETER 3.7 ASHRAE DESIGN HEAT LOSS. SQ FT PERCENTAGE DIFFERENCE 1 CONDUCTION DESIGN HEAT LOSS 1.7 OPERATIVE TEMPERATURE.3 6038.0 69.2 67.SQ FT 24.0 25.7 CEILING TEMPERATURE.0 5216.3 -7.5 2. BTU/HR. BTU/HR PERCENTAGE DIFFERENCE 3 ACTUAL HEAT INPUT.1088 0. BTU/HR PERCENTAGE DIFFERENCE 2 CONDUCTION DESIGN HEAT LOSS 2.0 -7.3 -7.3 6045. F 72.S.TABLE 35.7 6091. BTU/HR 6031.0 -7. DEG. DEG.9 69.7 191. BTU/HR PERCENTAGE:DIFFERENCE 4 F EFFECTIVE RADIANT FIELD.1 6548. HEATED FLOOR CASES .9 70.8 72.5 90. F 71.2 HEAT OUTPUT PER UNIT PANEL AREA.8 -7.T.0 84.0 69.9 27.2 6089.1 6548.8269 1. ONE WITH HALF GLASS FOR A 15' x 15' x 8' ROOM 83.2 INSIDE WALLS AND 2 OUTSIDE WALLS.5 5269.0 217.6 -20.1 67.0 MEAN RADIANT TEMPERATURE.9 60. DEG. DIMENSIONLESS 0. BTU/HR 6548. F .6 71.5 60.8 PERCENTAGE RADIATION 61.2 2.2 6089. DEC F PANEL AREA REQUIRED .0 '-7.1 PARAMETER 1.1 69.5 85.5 71.3 -19. SQ FT A.1 2.9 72.5 -18.4 6090.U. BTU/HR.9 -7. 5%.U .1. For this infrared base case and the base case for the U-tube types of infrared units discussed in Section 5. The other variables are as given in the base case.U .25 Btu/hr ft 2 F Ceiling . The U-factors were changed as follows: Walls . There is little effect in either situation on the difference between HLD and HLC. the follow. In these two situations the room is 15' x 15' x 8' and in each case one outside wall contains half glass.U = 0. cases for 3 inside walls and 2 inside walls are presented. The first situation was when there were no reflectors or deflectors on the units (which is not the normal operating condition) and the second is when there were reflectors or deflectors on the units and these reflectors are perfect and that the placement of the units is such that none of the direct radiation from the infrared units falls on the walls of the structure.0. Analysis has also been carried out for infrared modular (square or rectangular) heating types of units. Two situations were calculated for the modular and U-tube infrared cases.U = 0. This second situation would be the ideal design and placement case for infrared -109- .0 This change in wall and floor construction was made since these types of radiant units are most commonly applied to industrial buildings where the U-factors are commonly higher than what was specified in the original base case.9.8 .ing items were changed from the initial base case description. The ceiling height was set at nine feet and the outside design temperature at 3 F.25 Btu/hr ft2 F Floors . 5.Infrared Heating Cases.25 Btu/hr ft 2 F Glass Btu/hr ft2 F . In Tables 34 and 35.portion going from 61.7% to 55. This increase in design heating load is apparently due to a lower AUST because of increased U-factors and also more of the radiant energy being intercepted by the walls. The areas of the heaters which are . Figures 4 and 5 illustrate the first situation and Figure 27 illustrates the ideal situation with no direct radiation falling on the walls. This results in a greater design heat loss (up to 10% at 25 feet) than what is found from the ASHRAE standard design heat loss calculation. This shows up as an increased AUST (from 63 F at 9 feet to 66 F at 25 feet). Table 36 summarizes results for three infrared surface temperatures when there are no reflectors or deflectors. It should be kept in mind however that high temperature radiant units are normally mounted at the 12 to 15 feet level in an industrial building and use reflectors to direct the radiant energy away from the walls and toward the floor or occupants. shown in Table 36 are the total of 4 infrared heaters located at the ceiling (without reflectors) and were compared with several manufacturers and found to be in good agreement with their published ratings. In Table 37. results are presented for four 1700 F infrared units located at the ceiling (without reflectors or deflectors) for the base case as the ceiling height is extended to 25 feet. This lowering of the units and use of directive reflectors would nullify this 3 to 10% difference in design -110- . The percent difference betwen HLD and HLC was constant at approximately +3%. This indicates that as the heaters are raised in the room more of the radiant energy is absorbed by the larger wall area resulting in greater conduction losses. By looking at these two extremes --no reflectors or deflectors and perfect reflectors or deflectors the range of performance of generic units can be identified. This appeared to be the most reasonable approach to this type of heater since each manufacturer has a series of different reflector designs and suggestions or directions for placement of the units.modular and U-tube infrared heaters. &:•>:'•.•'/• pw-'S? p:---..^ rw:-' Vent System FIGURE 27. PLACEMENT OF INFRARED MODULAR UNITS WITH DEFLECTORS AND REFLECTORS TO PREVENT DIRECT WALL RADIATION TABLE 36. INFRARED MODULAR UNITS - BASE CASE WITH NO REFLECTORS OR DEFLECTORS AND VARIABLE SURFACE TEMPERATURE INFRA RED HEATER TEMPERATURE, D E C F 1600.0 1700.0 1800.0 2.4 2.0 1.6 ASHRAE DESIGN HEAT LOSS, BTU/HR 64378.8 64378.8 64378.8 ACTUAL DESIGN HEAT LOSS, BTU/HR 51343.2 51328.3 51315.4 -20.2 -20.3 -20.3 66224.4 66228.8 66232.6 ,'9 2.9 2.9 66224.4 66228.8 66232.6 2.9 2.9 2.9 66188.0 66198.6 66207.4 2.8 2.8 2.8 99.4 99.5 99.5 0.3 0.2 0.2 27832.4 33635.6 40312.8 FLOOR TEMPERATURE, DEG. F 73.7 73.7 73.7 ROOM AIR TEMPERATURE, DEG. F 60.4 60.4 60.4 MEAN RADIANT TEMPERATURE, DEG. F 84.7 84.7 84.8 OPERATIVE TEMPERATURE, DEG. F 74.0 74.0 74.0 EFFECTIVE RADIANT FIELD, BTU/HR. SQ FT 17.8 17.8 17.9 A.U.S.T, DEG. F 62.6 62.6 62.6 18.0769 .20.5157 23.1746 PANEL AREA REQUIRED , SQ FT PERCENTAGE DIFFERENCE 1 CONDUCTION DESIGN HEAT LOSS 1, BTU/HR PERCENTAGE DIFFERENCE 2 CONDUCTION DESIGN HEAT LOSS 2, BTU/HR PERCENTAGE DIFFERENCE 3 ACTUAL HEAT INPUT, BTU/HR PERCENTAGE DIFFERENCE 4 PERCENTAGE RADIATION PERCENT CEILING COVERED BY PANELS HEAT OUTPUT PER UNIT PANEL AREA, BTU/HR.SQ FT PARAMETER 1, BTU/HR.SQ FT.F PARAMETER 3, DIMENSIONLESS > 2 0.0006 0.0006 0.0005 TABLE 37. INFRARED MODULAR UNITS WITH NO 1700 F - EFFECT OF ROOM HEIGHT OR DEFLECTORS AND SURFACE TEMPERATURE AT ROOM HEIGHT, FT 9.0 10.0 12.0 15.0 20.0 25.0 PANEL AREA REQUIRED , SQ FT 2.0 2.1 2.3 2.6 3.3 4.0 ASHRAE DESIGN HEAT LOSS, BTU/HR 64378.8 67932.0 75038.4 85698.0 103464.0 121230.0 ACTUAL DESIGN HEAT LOSS, BTU/HR 51328.3 53973.1 59514.3 67341.4 83916.3 99536.9 -20.3 -20.5 -20.7 -21.4 -18.9 -17.9 66228.8 69934.7 77682.6 2.9 2.9 66228.8 69934.7 2.9 2.9 3.5 66198.6 69903.0 77647.5 2.8 2.9 3.5 3.4 8.0 10.2 99.5 99.5 99.5 99.5 99.5 99.5 0.2 0.2 0.3 0.3 0.4 0.4 33635.6 33646.9 33646.8 33648.2 33646.8 33646.2 FLOOR TEMPERATURE, DEG. F 73.7 73.5 73.4 72.4 73.7 73.3 ROOM AIR TEMPERATURE, DEG. F 60. 4 60.2 60.1 59.6 61.4 62.1 MEAN RADIANT TEMPERATURE, DEG. F 84.7 85.0 85.1 85.7 83.5 82.7 OPERATIVE TEMPERATURE, DEG. F 74.0 74.0 74.1 74.2 73.8 73.6 EFFECTIVE RADIANT FIELD, BTU/HR. SQ FT 17.8 18.2 18.3 19.2 16.2 15.1 A.U.S.T, DEG. F 62.6 62.6 62.8 62.7 65.1 66.3 20.5157 20.5179 20.5167 20.5109 20.5327 20.5414 0.0006 0.0006 0.0006 0.0006 0.0006 0.0006 PERCENTAGE DIFFERENCE 1 CONDUCTION DESIGN HEAT LOSS 1, BTU/HR PERCENTAGE DIFFERENCE 2 CONDUCTION DESIGN HEAT LOSS 2, BTU/HR PERCENTAGE DIFFERENCE 3 ACTUAL HEAT INPUT, BTU/HR PERCENTAGE DIFFERENCE 4 PERCENTAGE RADIATION PERCENT CEILING COVERED BY PANELS HEAT OUTPUT PER UNIT PANEL AREA, BTU/HR.SQ FT PARAMETER 1, BTU/HR.SQ FT.F PARAMETER 3, DIMENSIONLESS 3.5, 88684.3 111769.5 133661.2 3.5 8.0 10.3 77682.6- 88684.3 111769.5 133661.2 3.5 8.0 10.3 88644.7 111718.1 133599.0 TABLE 38. INFRARED MODULAR UNITS WITH NO REFLECTORS OR DEFLECTORS AND SURFACE TEMPERATURE AT 1700 F AND HEIGHT = 15' - EFFECT OF INFILTRATION RATE AIR CHANGES PER HOUR, AC/H PANEL AREA REQUIRED , SQ FT • 1.0 2.0 3.0 4.0 2.9 3.2 3.6 4.0 ASHRAE DESIGN HEAT LOSS, BTU/HR 94446.0 111942.0 129438.0 146934.0 ACTUAL DESIGN HEAT LOSS, BTU/HR 71677.8 79646.7 86809.3 93065.8 -24.1 -28.9 -32.9 -36.7 PERCENTAGE DIFFERENCE 1 CONDUCTION DESIGN HEAT LOSS 1, BTU/HR PERCENTAGE DIFFERENCE 2 CONDUCTION DESIGN HEAT LOSS 2, BTU/HR PERCENTAGE DIFFERENCE 3 ACTUAL HEAT INPUT, BTU/HR 95968.8 1092p2.3 121735.1 133075.9 1.6 -2.4 -6.0 -9.4 95968.8 109202.3 121735.1 133075.9 1.6 -2.4 -6.0 -9.4 95926.2 109154.3 121680.9 133015.7 1.6 -2.5 -6.0 -9.5 99.5 99.5 99.5 99.5 0.3 0.4 33647.4 FLOOR TEMPERATURE, DEG. F PERCENTAGE DIFFERENCE 4 PERCENTAGE RADIATION 0.4 0.4 33646.5 33644.7 33644.1 73.7 76.0 78.5 80.9 ROOM AIR TEMPERATURE, DEG. F 57.6 54.2 51.3 48.6 MEAN RADIANT TEMPERATURE, DEG. F 88.0 92.1 96.2 100.0 OPERATIVE TEMPERATURE, DEG. F 74.6 75.4 76.4 77.4 EFFECTIVE RADIANT FIELD, BTU/HR. SQ FT 22.3 27.8 33.2 38.0 A.U.S.T, DEG. F 63.3 64.3 . 65.8 67.3 20.4863 20.4432 20.4057 20.3721 0.0006 0.0005 0.0005 0.0005 PERCENT CEILING COVERED BY PANELS HEAT OUTPUT PER UNIT PANEL AREA, BTU/HR.SQ FT PARAMETER 1, BTU/HR.SQ FT.F PARAMETER 3, DIMENSIONLESS . heat loss as illustrated In the following calculations. In Table 38, the room was 15' highland the modular infrared units were at 1700 F and they did not have reflectors or deflectors. In this case (Table 38), the infiltration rate was changed between 1 and 4 ACH. As can be seen in Tables 38 and 37, the percent difference in HLD and HLC goes from +3.5% at 0.5 ACH to -9.4% at 4 ACH. This illustrates how much the infiltration rate affects the design load calculation. This is due to the exchange of lower temperature air for radiant systems when compared to forced air systems. With this increase in the infiltration rate, the floor temperature has an increase to 80.9 F, the room air temperature for comfort has decreased to 49 F and the mean radiant temperature has increased to 100 F. In this situation, the 4 ACH would most likely be beyond any normal situation (except for something such as spot heating) and does not represent a realistic situation. However, the importance of the change in the design heat loss load compared to the standard ASHRAE design load as infiltration is changed is strongly supported. In Table 39, the convection coefficient at the modular infrared units was changed by up to a factor of 5 for a 15' high room with 3 ACH and a modular infrared heater surface at 1700 F without reflectors or deflectors. As seen in this table, the assumption concerning the convection coefficient off of the heater surface has negligible effect on the calculations made here. If reflectors are used, there might be more of an effect due to more area available for convection heat transfer, however, it is expected to be negligible also. In Tables 40, 41, and 42 cases were run where the modular infrared units had perfect reflectors or deflectors and were positioned such that none of their direct radiation fell on the walls. This situation is illustrated in Figure 27 for an individual application. The situation in Table 40 is identical to the situation reported in Table 36 -- infrared base case -- except that Table 40 uses perfect reflectors and -115- TABLE 39. INFRARED MODULAR UNITS WITH NO REFLECTORS OR DEFLECTORS AND SURFACE TEMPERATURE AT 1700 F, 15' HIGH, 3 ACH - EFFECT OF CONVECTION COEFFICIENT MULTIPLIER CONVECTION MULTIPLIER 1.0 2.0 5.0 PANEL AREA REQUIRED , SQ FT 3.6 3.6 3.5 ASHRAE DESIGN HEAT LOSS, BTU/HR ACTUAL DESIGN HEAT LOSS, BTU/HR PERCENTAGE DIFFERENCE 1 CONDUCTION DESIGN HEAT LOSS 1, BTU/HR PERCENTAGE DIFFERENCE 2 CONDUCTION DESIGN HEAT LOSS 2, BTU/HR PERCENTAGE DIFFERENCE 3 ACTUAL HEAT INPUT, BTU/HR 129438.0 129438.0 129438.0 86809.3 87104.7 87964.9 -32.9 -32.7 -32.0 121735.1 121714.3 121664.4 -6.0 ' -6.0 -6.0 121735.1 121714.3 121664.4 -6.0 -6.0 -6.0 121680.9 121660.5 121611.7 PERCENTAGE DIFFERENCE 4 -6.0 -6.0 -6.0 PERCENTAGE RADIATION 99.5 98.9 97.3 0.4 0.4 0.4 33644.7 33830.2 34381.6 FLOOR TEMPERATURE, DEG. F 78.5 78.4 78.0 ROOM AIR TEMPERATURE, DEG. F 51.3 51.5 51.9 MEAN RADIANT TEMPERATURE, DEG. F 96.2 96.0 95.3 OPERATIVE TEMPERATURE, DEG. F 76.4 76.4 76.2 EFFECTIVE RADIANT FIELD, BTU/HR. SQ FT 33.2 32.9 32.0 A.U.S.T, DEG. F 65.8 65.7 65.5 20.4057 20.5200 20.8606 0.0005 0.0006 0.0006 PERCENT CEILING COVERED BY PANELS HEAT OUTPUT PER UNIT PANEL AREA, BTU/HR.SQ FT PARAMETER 1, BTU/HR.SQ FT.F PARAMETER 3,.DIMENSIONLESS 0 EFFECTIVE RADIANT FIELD. F 83. F .0006 0.U. 61. 0 63716.0 -1. BTU/HR. BTU/HR. DIMENSIONLESS 0.5 51177.1 18. SQ FT BTU/HR PERCENTAGE DIFFERENCE 1 CONDUCTION DESIGN HEAT LOSS 1 . 6 ASHRAE DESIGN HEAT LOSS. 8 ACTUAL DESIGN HEAT LOSS.6 63758. 0 PERCENTAGE RADIATION 99.4 99.T. 4 63755. BTU/HR.8 64378.-ja.0 -1.0 -1.4 63755. 2 -1. 8 -20.0661 20. DEG. DEG.2 60.9 84.00015 INFRA RED HEATER TEMPERATURE. BTU/HR PERCENT CEILING COVERED BY PANELS HEAT OUTPUT PER UNIT PANEL AREA.9 85.1 18.4 61. 5 63752. F 74. r INFRARED MODULAR UNITS WITH PERFECT REFLECTORS I N A 9 ' HIGH BASE CASE ROOM .KSfiiJKSS <P-i-*.0 74. SQ FT 18.SQ FT.F 18.4 61. F 60.0 33621.3 83. BTU/HR PERCENTAGE DIFFERENCE 2 CONDUCTION DESIGN HEAT LOSS 2 .9 1.EFFECT OF SURFACE TEMPERATURE 1600.SQ FT DEG.0 -1. 0 OPERATIVE TEMPERATURE. F 84. D E C F PANEL AREA REQUIRED .3 83.5 99.0006 0. 2 -1. 1 A. 2 MEAN RADIANT TEMPERATURE.S.2 0.0 -1.0 74. 3 ROOM AIR TEMPERATURE.7 40298.5051 23. 5 0.2 63733.5 -20. BTU/HR PERCENTAGE DIFFERENCE 3 ACTUAL HEAT INPUT. 0 FLOOR TEMPERATURE.5 -20.0 1800.4 51166.<s f^vi.0 1700.3 1. 51189.8 64378. 2 27819. 0 63752. 0 2.0 -1.3 0. 7 PERCENTAGE DIFFERENCE 4 -1. 4 PARAMETER 1.6 63758.164 1 PARAMETER 3.9 63726. BTU/HR 64378. DEG. DEG.2 60.. 0 58532.4 74.5 59.0 3.2 0.0 -22.3 84.0 -1. DEG.0006 .3 -3.4 53685.0 103464.1 60. DEG.2 2.0006 PERCENT CEILING COVERED BY PANELS HEAT OUTPUT PER UNIT PANEL AREA.4 3.7 -20.9 59.7 PERCENTAGE RADIATION 99.2 100011.8 60.6 66870. F 79.9 2.4 61.1 18.5 ROOM AIR TEMPERATURE. F 74.5 ACH AND 1700 F HEATERS .5 99.7 63755.4811 0.2 0.8 20. F 59.1 80311.2 FLOOR TEMPERATURE.2 0.5 EFFECTIVE RADIANT FIELD.7 -4.4 73314.1 -1.5 82005.0 25. DEG. SQ FT 20.5 58.4886 20.3 74.9 PERCENTAGE DIFFERENCE 1 CONDUCTION DESIGN HEAT LOSS 1.7 81969.3 -4.4818 20.7 60.4 33626.5 99.8 91.4 -21.5 OPERATIVE TEMPERATURE.4 83.3 -4.6 -3.8 -1.8 59. FT 8.7 87.S.0006 0.5 87.0 ACTUAL DESIGN HEAT LOSS.1 MEAN RADIANT TEMPERATURE. DEG.U.5 20.5 -21.0006 0.4 -3.5005 0. BTU/HR PERCENTAGE DIFFERENCE 3 ACTUAL HEAT INPUT.6 64378.7 -1.3 -23.2 33624. DEG.0 2.6 A.7 33606. BTU/HR. F 86.1 74.9 85. BTU/HR 60825.8 33621.2 86.5 101.8 67932.3 -4.4 ASHRAE DESIGN HEAT LOSS.5051 20. BTU/HR.5 92743. INFRARED MODULAR UNITS WITH PERFECT OF ROOM HEIGHT WITH 0. SQ FT" 1.8 84.3 0. BTU/HR .7 -4.4 58866.7 115477.0 10. BTU/HR PERCENTAGE DIFFERENCE 2 CONDUCTION DESIGN HEAT LOSS 2. BTU/HR.5 99. -22.0 74.6 100056.4 74. F 58.6 18.7 -1.0 12.6 63755.5 99.0006 0.0 -1.4 85698.1 -3.0006 0.8 20.3 66870.2 66839.3 21.2 108.EFFECT ROOM HEIGHT.6 -2.2 0.0 58558.6 100056.T.4 -23.3 21.2 51177. DIMENSIONLESS 33628.4975 20.2 115527.6 -2.0006 33614. BTU/HR 47042.5 99.2 59.5 0.SQ FT.4668 0.0 PANEL AREA REQUIRED .F PARAMETER 3.4 73314.3 -3.3 58.6 -2.7 63726.3 85.0 20.0 9.SQ FT PARAMETER 1.7 33621.7 1.7 82005.2 74.3 0.5 99.2 20.8 73280.3 -4.2 65878.8 87.9 58.2 115527.TABLE 41.0 75038.7 61.5 PERCENTAGE DIFFERENCE 4 -3.7 58558.0 121230.7 19.0 15. 6 20.1 42.1 112335.2 63. DEG.psa r~"? r-^ TABLE 42.6 ASHRAE DESIGN HEAT LOSS.1 100736.8 -6. F 89.7 112283.1 88056.7 49.3 -16.1 64.4 108.8 ROOM AIR TEMPERATURE.9 PERCENTAGE DIFFERENCE 4 • -6.0005 0.9 75. BTU/HR.0 -13.0 -13.SQ FT. F 56.0 111942. BTU/HR PERCENTAGE DIFFERENCE 2 CONDUCTION DESIGN HEAT LOSS 2. BTU/HR. DEG.2 -16.5 99. FT 2.3 100781.0 2.3 52.S.9 122257.9 82915.0 3. BTU/HR PERCENTAGE DIFFERENCE 3 so I IN 15' HIGH ROOM WITH HEATERS AT 1700 F - -26.1 MEAN RADIANT TEMPERATURE.3 -10. F 74.3 100781.5 30.4 0.0 PANEL AREA REQUIRED .1 PERCENTAGE DIFFERENCE 1 CONDUCTION DESIGN HEAT LOSS 1.0 129438.3592 20.2 PERCENTAGE RADIATION PERCENT CEILING COVERED BY PANELS HEAT OUTPUT PER UNIT PANEL AREA. DIMENSIONLESS 20.3196 PARAMETER 1.0 122314. BTU/HR 94446.F PARAMETER 3.9 -40.0005 0.3 EFFECTIVE RADIANT FIELD.1 99. F 60.3 3.4541 0.8 -6.SQ FT FLOOR TEMPERATURE. DEG.2 61.0005 .8 ACTUAL HEAT INPUT.0 -35.0 146934.5 A. DEG. SQ FT 24.8 88556.4064 20.2 78.8 99.6 3. INFRARED MODULAR UNITS WITH PERFECT REFLE( EFFECT OF INFILTRATION INFILTRATION RATE.1 46.2 -10.5 99.4 33622. BTU/HR.3 103.0 ACTUAL DESIGN HEAT LOSS. BTU/HR 88517.0 88556.2 104.3 99.3 0.8 33612.0 -13.0 4.5 0.U.2 -10.9 77.4 OPERATIVE TEMPERATURE.6 94.T.3 0.6 37.4 33607. AC/H 1.0 122314.0005 0.2 -16.2 33617.4 77235. BTU/HR 69909.9 -31.1 112335. F 94. DEG.5 99.0 3. SO. In these two tables. Also note in Tables 36 and 40 that the floor temperature has risen 10 F.unit placement. furniture and people will absorb this radiant energy and intercept it before it reaches the floor. Also note in Tables 37 and 41 that the floor temperature can get to too high of a value in the ideal situation (108 F)' but in the actual situation this will not be realized since equipment. Also observe that in Table 41 the air temperature for comfort is lower by up to 40 F. the floor temperature has increased theoretically by up to 25 F by use of reflectors and proper unit -120- . In these two tables the effect due to air infiltration is It is seen that the percent difference between HLD and HLC has changed from up to -9. It is now seen that the percent difference between HLD and HLC has gone from +3% to -1% indicating that proper placement of the infrared heaters can account for 4% in design load at these conditions. The situation in Table 41 is identical to the situation reported in Table 37 except that Table 41 uses perfect reflectors and placement of the heating units. Again. and the AUST is reduced by up to 7 F. The configuration and conditions in Table 42 are identical to those considered in Table 38 except that Table 42 considers ideal reflectors and unit heater placement. the effect of room height is considered. It is seen that the percent difference between HLD and HLC has changed from +10% with no reflectors to -5% with ideal reflectors and placement indicating that proper placement of the infrared heaters can account for 15% reduction in the design heat loss value. the room air temperature for comfort and MRT have not changed. considered.4% with no reflectors to up to -16. This indicates that with proper reflector design and unit placement of infrared heaters that up to 7% of the design heating load can be saved when considering height of the room. and the AUST has dropped about 1 F. the MRT is increased by up to 5 F.8% with ideal reflectors. These units have different orifice sizes for the same size and length of tube so that the units will operate at different average surface temperatures.8. the ceiling was extended in steps up to 20 feet. Table 45 is the same configuration and conditions as in Table 43 except that ideal reflectors and unit placement is -121- . The area calculated for the average surface temperature of 700 and 750 F agreed reasonably well with those presented by a manufacturer of these types units. In fact.U-Tube Infrared Cases. The results from these calculations are shown in Tables 43 thru 47. however. two situations were considered. Again. there is very little change in the results and the trends are similar.9 . the MRT increased by 3 F and the AUST reduced by 2 F. It is interesting to note here that the line source of radiation causes the opposite trend in the difference between HLD and HLC than was observed for the modular infrared units.8 for the infrared modular units except for this case two U-tubes were used in the space instead of the four modular units. The results given in Table 44 are for a 750 F average surface temperature. 5. In Table 43.placement. no reflectors and then with ideal reflectors and placement. . Tables 43 and 44 give results for the U-tube heaters that do not have reflectors or deflectors. The behavior of the U-tube units appears to be similar to that of the ceiling panel heating types of units. the average surface temperature was varied between 700 and 900 F. Tables 45-47 present the results for the U-tube infrared units which have ideal reflectors and unit placement. The same conclusions can be drawn from these results as from the infrared modular results discussed in Section 5. The same base case parameters were used as described in Section 5. The U-tube or straight tube type of configuration for vented gas-fired infrared units were also analyzed. The room air temperature for comfort has been reduced by 2 F. BTU/HR.0 MEAN RADIANT TEMPERATURE.1 17.0 2. F 60.5 66317.2 84.8 60.T.2 84.U-TUBE INFRARED UNITS . SO.2 98.9 2.5 3.0012 : . DEG.0013 0. BTU/HR PERCENTAGE DIFFERENCE 3 ACTUAL HEAT INPUT.5 3.1 OPERATIVE TEMPERATURE.1 62.9 66649.U.4 1.6 3834.5 51791.1 3.1 5.2 84.3 2728.SQ FT PARAMETER 1.6 -19.2702 0.2 73.5 98.1 97. DEG.3 73.3 3.5 U-TUBE IR-HEATER TEMPERATURE.7 66423. BTU/HR 51692.9 66640. BTU/HR.5 3.9 73.5 3.SQ FT.0012 0.8 3.CHANGE IN TUBE SURFACE 900.8 64378.7 -19.2 3.5 66375. F 84.S. BTU/HR PERCENTAGE DIFFERENCE 2 CONDUCTION DESIGN HEAT LOSS 2.8 64378. F 62.8 17.F 6.3 73.3 66630. DEG.1 51831.2 17.0014 PERCENTAGE DIFFERENCE 4 PERCENTAGE RADIATION PERCENT CEILING COVERED BY PANELS HEAT OUTPUT PER UNIT PANEL AREA. BTU/HR 66464.3 ASHRAE DESIGN HEAT LOSS.9 73.1883 4.BASE CASE WITH NO TEMPERATURE OR DEFLECTORS .7132 4.3 3247.0011 0.2 3.9 1.1 84.0 66619.0 850.1 17.3 2.1 ROOM AIR TEMPERATURE.6 66247. DIMENSIONLESS 0. F 73.5 3.5 66658.7 14. BTU/HR. EFFECTIVE RADIANT FIELD.6962 PARAMETER 3.9 -19.2 73. D E C F PANEL AREA REQUIRED . SQ FT 17.7 5234. F 73.0 700.0 66619. DEG.2386 5.9 A.0 12.6 -19.0 16.8 64378.9 .8 64378.4 98.9 73.9 60.1 62.7 -19. DEG.3 3. FT PERCENTAGE DIFFERENCE 1 CONDUCTION DESIGN HEAT LOSS 1.0 800.9 61.5 3.8 60.9 66649.0 750.8 4494.3 3.9 98.9 66630.9 73.3 20.4 24.1 62.1 51721.5 FLOOR TEMPERATURE.8 ACTUAL DESIGN HEAT LOSS.5 66658.5 3.1 62.6 1.6 51754. BTU/HR 64378.9 66640. 0013 0.1 60.5 69.2 OPERATIVE TEMPERATURE.0 PANEL AREA REQUIRED .SQ FT.9 74.3 65300. PERCENTAGE DIFFERENCE 1 CONDUCTION DESIGN HEAT LOSS 1.7 61. DEG.2 74. DEG.4 21.7 76463.0 17.S.0 100683.8 -23.4 59. F 84. p'"i W~m Wr-'-'i »r*3 f-'-* f : TABLE 44.0013 ACTUAL HEAT INPUT.0 20.1 58.0 -19.4 23.4 -2.1 98.3 75864.4 2.6984 4.0013 0.3 3248.0 10.5 ROOM AIR TEMPERATURE.5 1.3 57.0 20. BTU/HR 51791.6 PARAMETER 1.U.6 75.3 30.1 -2. BTU/HR PERCENTAGE DIFFERENCE 2 CONDUCTION DESIGN HEAT LOSS 2.7 98. DEG.3 0.SQ FT 85632.4 3247.7" 76463.1 61.3 3.3 101112. SQ FT .7 66. F 60.9 25.0 •2.0 87. U-TUBE INFRARED UNITS .EFFECT 9. F 73.0 74.4 85698.1 98. BTU/HR.9 66317.9 1. DIMENSIONLESS 0.2 A.0 69924.9 55. BTU/HR PERCENTAGE DIFFERENCE 4 PERCENTAGE RADIATION PERCENT CEILING COVERED BY PANELS HEAT OUTPUT PER UNIT PANEL AREA.0 15.6 71.8 MEAN RADIANT TEMPERATURE.4 -0.9 19.5 66630.3 FLOOR TEMPERATURE.5 2.9! 1.7056 4.6 -20.9 3.4 -2.0013 0.7107 4.0013 0.4 26.8 -26.3 86012. DEG.0 103464.9 ASHRAE DESIGN HEAT LOSS. BTU/HR.6 21.1 98.2 3249.9 69924.0 12.8 67932.3 0.3 -21. BTU/HR.0 ACTUAL DESIGN HEAT LOSS.4 3.F 4.1 84.3 76115.0 EFFECTIVE RADIANT FIELD.3 101112.T.8 86.7 58689. BTU/HR 64378. F 62.W~m f S .5 . BTU/HR PERCENTAGE DIFFERENCE 3 3.3 2.0 2. F 73. SQ FT 17.9 60.1 98. FT AND NO REFLECTORS OR DEFLECTORS .1 54135.7 90.6884 PARAMETER 3.6 2.750 F SURFACE OF CHANGE IN ROOM HEIGHT ROOM HEIGHT.1 2.2 72.7 66630.5 3254.0 75038.7132 4. DEG.7 3251.6 69600.0 86012. BTU/HR 51414.4 98.4 61.8 64378.4 PARAMETER 1.CHANGE IN TUBE SURFACE TEMPERATURE 700.6 -1.1 -1.0 63655.8 5220.0013 0.8 ACTUAL DESIGN HEAT LOSS. DIMENS'I'ONLESS 0.0011 U.1 3233.2 1.1 51343. BTU/HR PERCENT CEILING COVERED BY PANELS HEAT OUTPUT PER UNIT PANEL AREA.1 4480.6898 5.7 63639.0 800.1 63618.9 98.2 ASHRAE DESIGN HEAT LOSS.0014 0.0 63682.7 FLOOR TEMPERATURE. U-TUBE INFRARED UNITS .3 83..9 63670.0 74.6 1.F 4. BTU/HR. DEG.4 -1. DEG. BTU/HR PERCENTAGE DIFFERENCE 3 ACTUAL HEAT INPUT.8 1.0 74.2 -20.8 17.6 16.7 84.0 EFFECTIVE RADIANT FIELD.4 2714.8 64378.6 84.2 -1.0 850.SQ FT.4 61.2189 PARAMETER 3.7 OPERATIVE TEMPERATURE.6757 6. DEG.1 51377.5 60.8 64378.7 17.3 19. BTU/HR.1 -1.TABLE 45. F 61. BTU/HR PERCENTAGE DIFFERENCE 2 CONDUCTION DESIGN HEAT LOSS 2.1 63256.8 63334. F 83.U.1662 5.1 -1.0 63682.TUBE HEATER TEMPERATURE.7 84.0 23.0 74.6 2. F 74.0 900.8 A.3 83.2 63493.9 63398.7 -1.3 ROOM AIR TEMPERATURE.3 83.7 63639.8 64378. DEG. BTU/HR 64378.1 51396.T.4 PERCENTAGE RADIATION 97. SQ FT PERCENTAGE DIFFERENCE 1 CONDUCTION DESIGN HEAT LOSS 1.SQ FT .4 60.1 -1.5 -1.9 63670. DEG.0 -1.2 63618.1 -1.2447 4. SQ FT 17.6 14.0012 0.1 -1.2 63450.0 -1.2 12.7 17. F PANEL AREA REQUIRED .6 84.S.9 51360.3 3820. DEG. F 84.BASE CASE WITH IDEAL REFLECTORS AND PLACEMENT . F 60.5 60.2 -20. BTU/HR.2 -20.4 61.0 63655.0012 0.7 17.5 2.0 74.2 98.0 750.1 98.4 61.2 -1.3 83.2 PERCENTAGE DIFFERENCE 4 -1.1 -20.5 60.0 -20.4 MEAN RADIANT TEMPERATURE. 1 60.6 25. F 60.1 2.3 EFFECTIVE RADIANT FIELD.7 18. DEG.0 3214.0 75038.5 OPERATIVE TEMPERATURE.1 -1.8 59.1 85.S.1 74. BTU/HR PERCENTAGE DIFFERENCE 2 CONDUCTION DESIGN HEAT LOSS 2.2 -1.8 72810.3 19. F 83.1 98. BTU/HR.1 59.0013 .9 -22.1 81802.0013 0.1 53903.3 99252. BTU/HR.4 85698.3 84.0 -4.4 61.2 2.0 74.5 59074.SQ FT PARAMETER 1. F 74. 6898 4.3 2.0 66742.7 -21.0 ACTUAL DESIGN HEAT LOSS.0 20.8 -2.0 19.5 -3. DEG.6 -2.0 86.5 2.6 66742. DIMENSIONLESS 0.2 A.1 ' 73156.6 99714.8 3.0013 0.5 -20.0013 0. FT PANEL AREA REQUIRED .6777 4.8 -2. BTU/HR.9 20.8 81425.5 80365. BTU/HR PERCENTAGE DIFFERENCE 3 63639.SQ FT.5 60.5 -4.0 20.U.4 ACTUAL HEAT INPUT.9 ASHRAE DESIGN HEAT LOSS. BTU/HR 64378.1 -1.0 -5.7 58.1 98.T.5 58. BTU/HR PERCENT CEILING COVERED BY PANELS HEAT OUTPUT PER UNIT PANEL AREA.0 15.6 99714.7 87.6 66060.0 PERCENTAGE DIFFERENCE 4 -1.6517 PARAMETER 3.F 4'.0 74. F 61. F 84.2 ROOM HEIGHT.9 MEAN RADIANT TEMPERATURE.6 22.3 -22. DEG.4 74.0 10. SQ FT PERCENTAGE DIFFERENCE 1 CONDUCTION DESIGN HEAT LOSS 1.9 91.1 98. SQ FT 17.6643 4. BTU/HR 51396.0 103464.6 20. DEG.2 -20.1 PERCENTAGE RADIATION 98.750 F SURFACE EFFECT OF CHANGE IN ROOM HEIGHT AND WITH IDEAL REFLECTORS AND PLACEMENT - 9.1 81802.7 101.9 66424.6 87.0 -1.1 73156.2 30.7 3225.1 98.0013 0.6852 4. DEG.3 FLOOR TEMPERATURE.5 -3.7 ROOM AIR TEMPERATURE.7 3228.3 3231.3 63639.8 67932.U-TUBE INFRARED UNITS .0 12.6 85.7 60.5 -4.2 -3.4 3233.6 63334. 1 62.7 -40. F 89.9 83180.2 3217.0 27.0 146934. DEG.5 -17.2 111454. F 60.6152 4.0 4.9 77.0013 0. SQ FT 24. BTU/HR 70097. BTU/HR 88345.2 -13.0 2.SQ FT.9 -17.1 98.0 98. BTU/HR 94446. DEG.0 87940.0 3.1 99. F 74.0 3. SQ FT AND PLACEMENT IN A 15' HIGH ROOM - 1.8 49.4 121984.2 78.4 PERCENTAGE RADIATION 98. DEG.5 -17.1 61.0013 0.8 -6.TABLE 47.0 88345.0 111942.3 OPERATIVE TEMPERATURE.0013 PERCENT CEILING COVERED BY PANELS HEAT OUTPUT PER UNIT PANEL AREA.5842 4.0 PERCENTAGE DIFFERENCE 1 CONDUCTION DESIGN HEAT LOSS 1.4 52.3 31.0 129438.6466 4.5577 PARAMETER 3.5 111983.0 3. BTU/HR.4 121393.U.2 MEAN RADIANT TEMPERATURE.6 104.4 ROOM AIR TEMPERATURE.F 4.5 111983. U-TUBE INFRARED UNITS AT 750°F WITH IDEAL EFFECT OF CHANGES IN AIR INFILTRATION INFILTRATION RATE.8 ASHRAE DESIGN HEAT LOSS.5 3.4 94.9 64.1 103.6 77439.1 34. BTU/HR PERCENTAGE DIFFERENCE 3 ACTUAL HEAT INPUT. BTU/HR PERCENTAGE DIFFERENCE 2 CONDUCTION DESIGN HEAT LOSS 2.5 -10.9 -10. DIMENSItONLESS 0.5 -10. DEG.5 PARAMETER 1.0013 0.8 109. DEG.8 -30. AC/H PANEL AREA REQUIRED .4 99. F 56.6 100498.0 98.T.SQ FT . BTU/HR.3 FLOOR TEMPERATURE.0 100038.8 -6.8 42.9 75.4 -25.4 A.3 3211.3 EFFECTIVE RADIANT FIELD.8 -35.6 100498.4 121984.1 PERCENTAGE DIFFERENCE 4 -6.0 ACTUAL DESIGN HEAT LOSS.9 4.4 36.1 88205.9 3207. BTU/HR.7 37.2 -13.3 46.S.2 30.2 3222.6 -13. F 94. at the same time the floor temperature becomes very high (up to 109 F in the theoretical undisturbed case) and the air temperature for comfort has been reduced to 46 F. This change results in a decrease in the percent difference between HLD and HLC of up to 5% for this set of conditions. In Table 46. The trend in the floor temperature has been changed in Table 46 because more area of heating surface has been installed and none of this heat is intercepted by the walls. ideal reflectors and unit placement can reduce the installed heating capacity up to 5%. This change results in a decrease in the percent difference between HLD and HLC of up to 4 1/2%' for 9' high ceilings. the results for U-tube infrared units "at-750 F with ideal reflectors and placement in a 15' high room are given for infiltration rates changing from 1 to 4 ACH. However. the same configuration and conditions as in Table 44 are considered except that in Table 46 ideal reflectors and unit placement are considered. This causes the floor temperature to approach 100 F for the 20' high room. This shows that the percent difference between HLD and HLC can be up to 17% by use of ideal reflectors. HLD is the ASHRAE standard heat loss calculation procedure and HLC is the design heat loss for a -127- . Also showing up in this calculation is an increase of about 10 F in the floor temperature when all of the infrared radiant energy from the heater is reflected directly to the floors with none impinging on the walls. Therefore. These are rather extreme situations and most likely would not be encountered in a total heat situation for infrared heaters. the use of ideal reflectors and unit placement can result in the savings of up to 5% in the design heat loss.Summary of Design Heating Calculations The results in the changes in percent difference between HLD and HLCG given in Tables 6 thru 47 have been summarized in Table 48. Again.10 .used in Table 45. In Table 47. 5. 0 C2 Base case with 0.0 to 4.0 Panels from 120 to 180oF -4.1 to 3. gradient per foot 8 to 25 ft. •7.75oF/ft.1 to -4.2 C3 Base case with loF temp. -7.0 ACH •1.5 to -3. gradient 1. gradient per foot 8 to 25 ft.94 -4.4 C7 25' High base case with 0.2 to +1.7 to -5. -6. temp.0 ACH +2. gradient and U-factors used in infrared cases PANEL HEATING PI P2 Base case Base case with panels at 140oF -128- . -7.2 -panel from 0.88 to 0.2 C5 15' High base case with 0.0 to 4.9 to +7.0 ACH -7.75oF/ft temp.3 1. gradient per foot 8 to 25 ft. gradient per foot 8 to 25 ft.9 to +16.5 C4 Base case with 1.75oF/ft temp.5oF temp.0 to 4.TABLE 48 Fixed Conditions SUMMARY OF CALCULATED RESULTS Variable being changed and its range Range of the difference in percent between HLCG and HLD Percentage Difference 3 (%) Forced Air Heating Cl Base case with 0.2 C6 25' High base case with 0. gradient 1.3 to -1.5 to +11.75oF temp.5oF temp. 6 P7 Base case w i t h p a n e l s a t 130<>F 0.2 P12 Base case with panels at 140oF Room size change from 15' x 30' to 100' x 100' -2.4 Pll Base case Panels from 130 to 170oF U increase from 30 to 100% -2.2 to -3.0 P10 P16 Perimeter panel.5 to 4.0 P13 Base case with panels at 140oF Room height changed from 8' to 25' -5.0 ACH -3.5 to -4.TABLE 48 (CONTINUED) P3 Base case with panels at 140oF e walls f r o m 0.7 P8 Base case with panels at 150oF 0.5 to -2.5 to -15.5 to 4. 3 inside walls.4 Panel from 175 to 185oF -0. 1 outside wall with glass -129- .5 P6 Base case with panel h c multiplied by ten Panels from 120 to I8O0F -4. 36" panel.0 ACH -3.4 to -4.0 ACH -3.8 to 0.5 to -15.7 to -5.7 P9 Base case with panels at 170oF 0.3 P15 Base case with panels at 140oF Number of panels changed from 1 to 6 -4.6 Base case with panels at 140oF Glass quantity from 0 to 50% of total outside wall area -4. 15' x 15' x 8' room.2 to -4.7 to -4.3 P5 Base case with panel h c multiplied by five Panels from 120 to I8O0F -4.1 to -0.95 •1.3 to 0.5 to -15.6 to -5.4 P14 Base case with panels at 130oF Outside design temperature changed from -5 to 15oF -3.4 P4 Base case with panel h c doubled Panels from 120 to 180oF -4.7 to -6.5 to 4. 15' height.0 ACH h c from heater changed up to a factor of 5 -130- +2.Infrared.TABLE 48 (CONTINUED) HEATED FLOOR Fl Base case-30'x30'x8' Floor temperature from 81 to 85oF F2 Base case with floor at 84oF Outside design temperature from 5 to 20oF -6. 15' height.0 to 4.9 to +10. 2 inside walls and 2 outside walls with halfglass Floor temperature from 83 to 85oF -7.9.15 -6. No reflectors/ deflectors Room height changed from 9' to 25' F3 F4 F5 F6 -6.4 -6. No reflectors /deflectors Base case .0 . 3 inside walls and one outside wall with half glass Floor temperature from 83 to 85oF -5. 1700oF surface temperature. 1700oF surface temperature. 3 ACH. No reflectors/ deflectors Surface temperature from 1600 to 1800oF +2.9 Base case with floor at 84oF 0.25 ACH •6.1 Base case with floor at 84oF Floor U-factor changed from 0.0 Base case . No reflectors/deflectors 1.9 to -7.0 to -7.0 INFRARED MODULAR UNITS II 12 13 14 Base case .9 Base case with 15' x 15' x 8'.Infrared.6 to .9 to +2.4 Base case with 15' x 15' x 8'.07 to 0.3 1.Infrared.Infrared.9 to -5.9 to -7.9 to -7.5 to 1.5 to -8.9 Base case . 1700oF surface temperature. 3 U3 Base case .0 1. 9' height. 1/2 ACH with reflectors/ deflectors (CONTINUED) Surface temperature from 1600 to I8OO0F -1.5 to -17.0 ACH •6. with reflectors/deflectors Base case .. 15' height.2 to -1. 1700oF.Infrared with reflectors/ Tube surface changed from 700oF to 900oF •1.5 to +3.0 to 4.0 to 4. 750°F tube temp. 750oF tube temp.TABLE 48 15 Base case .5 to -2.0 ik: m -131- .2 to -16.Infrared.Infrared 750oF tube surface.7 to -4.Infrared.U-TUBES tLi Ul Base case .8 Room height 8' to 25' •3.5 U2 Base case . No reflectors/ deflectors Room height changed from 9' to 20' +3.Infrared No reflectors/ deflectors Tube surface changed from 900oF to 700oF +3. 15' height with reflectors/ deflectors 1. with reflectors/ deflectors 17 Base case .Infrared.3 U5 Base case .7 16 INFRARED .1 U4 Base case .1 to -3.0 to -1.Infrared. 1700oF. with reflectors/ deflectors Room height changed from 9' to 20' •1.Infrared.0 ACH -6. Ul . -132- . A listing of the computer program which was used to perform all of these calculations is given in APPENDIX-C.1 7 . and Infrared U-Tubes . For some situations (C6 and C7) the ASHRAE standard procedure can undersize the system by up to 15%. This summary in Table 48 shows that the ASHRAE design heat loss calculation can oversize a system up to about 17% but the most common value is 4 to 7% oversizing for all of the variables and conditions considered here. A list of the input variables is also given there. Each of the types of heating systems are identified such as: Forced Air Heating . Heated Floor .PI .P16.Fl . Infrared Modular Units . Table 48 describes the basic cases from the previous tables.F6.1 1 . Panel Heating .C7.U5.space considering the actual conduction through the -walls and the infiltration load based on the air change method. gives the variable being changed and its range of change.CI . and the percent difference in the two design heat loss calculations. 5 ACH.0 .DESIGN PROCEDURES In the previous section. the oversizing can be up to 17%. room size. The variation shown in Figure 28 is recommended as the only reduction factor to use for sizing radiant heating systems. However.6. it was shown in Table 48 that the use of the ASHRAE Design Heating Load Procedure (HLD) [1] would result typically in a slightly oversized heating system.0 ACH. However. this oversizing is up to 5% at 0.0 ACH it is 9 1/2%. If the infiltration rate is at 4 ACH.0 ACH it is 13% and at 4.0 .0 ACH it is 5 1/2%. A single line has been drawn through the data to represent all four types of radiant heating systems such that at 0.5 ACH. for modular and U-tube infrared (high and medium temperature respectively) units with good reflectors and proper location such that no direct infrared radiation impinges on the walls.0 ACH it is 16%. as the infiltration rate increases and a temperature gradient is present the HLD procedure can underestimate the required heater -133- . In Figure 28.CALCULATION OF DESIGN HEATING LOADS. It should be pointed out that the ASHRAE Standard Design Heating Load procedure (HLD) also does some overestimating (up to 7%) for forced air heating systems. the percent reduction of standard design load is plotted against the air infiltration rate for panel heating. 5.P16) that for variations in panel temperature.5 ACH the reduction in heating unit size is 4%. Examination of Table 48 shows for panel heating systems (P-l . at 1.5 ACH it is 7 1/2%. at 1. and room height that this oversizing is about 3-6% for 0.5 ACH. The conclusion of this investigation is that the air infiltration rate is the only variable in the design heat loss calculation which affects in a meaningful way the results for the sizing of radiant heating units. at 2. Likewise. In a similar way for the heated floor situation the oversizing is about 7% at 0. for larger infiltration rates this oversizing can be up to 15% for 4. at 3. heated floor and infrared modular and U-tube units-. 750°F 14 M W Q 12 H CO r 10 b O O M H O & P W H 25 W u w AIR CHANGES PER HOUR FIGURE 28. PERCENT REDUCTION OF STANDARD DESIGN LOAD AS A FUNCTION OF AIR INFILTRATION RATE . High Temperature Unit. Medium Temperature Tubes. 170 °F Infrared.20 18 ~ • i • r Panels at 130. 150. 1700°F ' 16 o g Infrared. APPENDIX D contains a reproduction of Chapter 8 from the 1984 ASHRAE Systems Handbook. There is not any overwhelming evidence to reduce the design heat loss values for any of the other parameters in the radiant heating situation. The following are the recommended design steps for panel heating systems. Some additional recommendations are also added. From these results.the design procedures.1 . warm water panels with embedded pipe (plaster ceiling and concrete ceiling).Radiant Ceiling Panel Heating Systems In Chapter 8 of the 1984 ASHRAE Systems Handbook [2]. The only change from what appears in Chapter 8 of the 1984 Systems Handboook is that the design heat loss is reduced as a function of infiltration rate as given in Figure 28. The current design steps and procedures given in Chapter 8 of 1984 Systems have been reviewed and compared with the original work in the literature which is presented in APPENDIX B . it is recommended for radiant heating systems that the ASHRAE design heat loss calculation procedure presented in Chapter 25 of the Handbook of Fundamentals [1] be used with a reduction in the final value made according to the estimated infiltration rate as presented in Figure 28. 6. various manufacturers have used these procedures for many years and have not reported any deficiencies in.ANNOTATED BIBLIOGRAPHY. There were no serious problems or difficulties with the assumptions made in compiling the curves and tables used in the procedures. The following procedures are suggested for designing radiant heating and cooling systems. Calculate the hourly rate of heat loss for each room using the procedures given in Chapter 25 of the 1985 ASHRAE Handbook of -135- . In addition. Panel HeatingrSystem Design Steps 1. and electric ceiling panels.size by up to 16% for a 25' high room. there is a section on "Panel Heating System Design" and this includes data and examples for metal ceiling panels. The calculation procedure was not able to calculate the actual outside wall or glass temperature for summer conditions since it did not consider solar effects on the wall or glass. 8. 6. 5. 2. 9. Select insulation for the reverse side and edge of panel 7. 4. The following general rules should be followed: 1) Place panels near cold areas where the heat losses occur. the effect of each assumption or choice on comfort should be considered carefully. 3.5. Determine the required panel surface temperature.Radiant Ceiling Panel Cooling Systems The computer procedure which was developed and discussed in SECTION . Determine any other temperatures that are required. Determine panel heat loss and required input to the panel. The procedure was developed basically for heating design load calculations where solar effects would not be considered -136- . 2) Do not use high temperature ceiling panels in very low ceilings.2 . Design the system for heating the panels according to conventional practice.0 and presented inAPPENDIX-C has been applied to several cases of radiant panel cooling. Select the means of heating the panel and the size and location of the heating elements. 6. Specific design examples are given in Chapter 8 of the 1984 ASHRAE Systems Handbook which appears in APPENDIX D. Determine the available area for panels in each room. Reduce this heat loss by the amount given in Figure 28 for the specific estimated air infiltration rates.Fundamentals fll. 3) Keep floor temperatures at or below 85 F (29 C). Always consider the manufacturers recommendations for pre-engineered heating panel systems. Calculate the required unit panel output. In the design steps. The analysis for the procedure for sizing radiant panel cooling systems involved examination of the original ASHRAE research work and the procedure for panel cooling given in Chapter 8 of the 1984 ASHRAE Systems Handbook. The procedure is as follows. the correction to the design load given in Figure 28 does not apply for the design cooling load. the infiltration load for summer design is expected to be significantly less since the inside .outside air temperature difference (stack effect) is much smaller and the summer design wind velocity (typically 7 1/2 mph) is typically half of the winter design wind velocity (15 mph) . -137- . Also. It is recommended that the ASHRAE design cooling load procedures (for commercial buildings and residential buildings) presented in Chapter 26 of the 1985 ASHRAE Handbook of Fundamentals" [1] be used directly. In addition. This appears to be sufficient for the cooling situation since in that case the portion of sensible heat removed by radiation is significantly less since surface temperature differences are less in the cooling mode than in the heating mode. the radiant cooling system is not able to absorb the latent load so that the ventilation air brought to the cooling space absorbs this latent load and at the same time absorbs some of the sensible load. Also. there has not been any research located since the ASHRAE work in the 1950's that would invalidate the current design procedures. The procedures given in Chapter 8 of the 1984 ASHRAE Systems Handbook [2] were checked and verified with the references and original work done by ASHRAE. In addition.at the design time. several manufacturers have been using this procedure for many years and have not reported any deficiencies in the procedure. For these reasons. There are no engineering and/or comfort reasons to expect a reduction in the design cooling load calculations. 6. In the evaluation of this design procedure. 7. 14. Calculate room heat loss using the procedures given in Chapter 25 of the 1985 Handbook of Fundamentals f 11 . 5. 3. it is seen that the location of the air diffusers relative to the panel sections does not enter into the -138- . Establish minimum supply air quatity. 13. Select mean water temperature for cooling. Calculate the sensible cooling available from the air. Determine the room design dry-bulb temperature. 10. Determine panel area for heating. 4. Calculate room sensible and latent heat gains using the ASHRAE procedure given in Chapter 26 in the 1985 ASHRAE Handbook of Fundamentals. design for the heating situation. 8.PANEL COOLING SYSTEM DESIGN 1. Now. Specific design examples are given in Chapter 8 of the 1984 ASHRAE Systems Handbook which is reproduced in APPENDIX D. 12. relative humidity and dewpoint. Calculate the latent cooling available from the air. Determine water flow rate and pressure drop. Always consider the manufacturers recommendations for placement. Determine panel cooling load. 9. Design the panel arrangement. 2. Select mean water temperature for heating. Determine required panel area. Reduce this heat loss by the amount given in Figure 28 for the specific estimated air infiltration rate. sizing and insulation of pre-engineered panel heating and cooling systems. 11. D. it should be emphasized that the latent heat gain must be absorbed by an independent source. For typical design situations. this should not alter the It should also be pointed out that the lighting load for the cooling case should be carefully evaluated since it will be a major contributor to the cooling load. The source specified in the design procedure is the ventilation air which is dehumidified separately. 6. design procedure. we are only interested in the design heat loss calculation and the placement of the heating surfaces. The design procedure does not account for different types of lighting fixtures and these loads should be incorporated into the design heat gain calculation. tube spacing and lower insulation) in the design process and their effect on the upward heat delivered by the system. and electric floor slab heating) is presented in Chapter 8 of the 1984 ASHRAE Systems Handbook T21.3 . It would also be possible to operate various types of dehumidi- fiers within the space to absorb this latent load or to use chemical dehumidification. It would be expected that if the air is diffused in close proximity to the ceiling panels that the cooling and heating performance would be altered slightly. The actual design process for radiant systems was not part of this project. There were some papers obtained in the literature search which discussed the physical parameters (slab thickness. The only modification to be made to that procedure is the reduction in the design heat loss calculation shown in Figure 28 for high infiltration air changes per hour.design procedure. Here.Heated Floor Systems A design procedure for heated floors (concrete floor panels for slab-ongrade. There are several design - examples given in Chapter 8 of the 1984 ASHRAE Systems Handbook f21 which is included in APPENDIX . Also.1. -139- . as indicated in SECTION 6. concrete floor panels for intermediate slabs. Lienhard and Tezduyar [22] in an UNPUBLISHED private report in 1985. since this is an unpublished report it is not a valid source of information for changing -140- .Grammling [19] in a 1985 ASHRAE paper pointed out that in the German standard DIN 4725. methods were developed for testing the thermal performance of hydronic floor-heating systems. In addition. Most of the systems have been tested with the so-called plate apparatus. These design recommendations are conservative because both the downward and edgewise heat loss and panel thermal resistance are over estimated. Some similar conclusions were presented by Shamsundar. The ASHRAE panel heating model does not represent the panel heat loss mechanisms correctly but the design recommendations are adequate and slightly conservative for designing both bare and covered radiant floorheating panels with no infiltration and an AUST equal to the room air temperature. However. These conclusions were also presented in an ASHRAE paper by Hogan and Blackwell [21]. It is clear from these results that exact performance measurements under controlled thermal conditions are necessary for designing and laying out unique floor-heating systems. numerous measurements have been made for these types of heating systems. This was done using a steady state and a transient numerical model of the heated floor slab. The results of the tests show that measured values of performance differ significantly from figures published in the literature or company catalogs. This is most likely why it has not been detected in existing designs. Hogan [20] in a MS thesis reviewed and evaluated the ASHRAE design recommendations given in Chapter 8 of the 1984 ASHRAE Systems Handbook f21. They have shown that the ASHRAE procedure is erroneous (using numerical simulation) and that it can be modified to make it more correct. Some of the error that they note is on the conservative side and some of it is underestimating the requirements so that the errors appear to cancel each other for most conditions. combustible or potentially toxic vapors in the building and restrictions for moisture level requirements. They cover both gas and electric as well as various intensity levels (porous refractory. with minor variations between manufacturers. They indicate that the ASHRAE procedure can also be used for systems with plastic pipe by using simple multipliers for various pipe diameters. radiant tube. They have specific types of reflectors and are mounted at various angles so as to prevent radiating the wall. quartz tube.Hiph and Medium Temperature Infrared Systems The design guidelines provided by manufacturers of infrared heating systems (14 were made available) have been reviewed. and metal sheath electric). -141- . and a lower air temperature required for comfort when radiant energy is used for heating. usage schedules. At some point in the design of the unvented gas units. These begin with a heating survey taking note of building materials. The size of the heaters are then selected from the manufacturers published data.4 .the current design procedure. lateral spacing of heaters. 6. Providing the necessary make-up air and exhaust for these systems is an extremely important consideration. These units are usually mounted along the perimeter where the high heat losses occur. A standard ASHRAE heat loss calculation is suggested along with a reduction recommendation ranging from 0 to 25% with the usual value being about 15%. less heat loss due to a reduction in air temperature stratification in the building. consideration of minimum dilution air to control the CO2 and condensation possibilities must be taken into account. Various reasons are given for this reduction in design heat load: unvented gas fired units are more efficient. The design guidelines for all of these units are very similar. and distance from combustible materials. Specific details are given concerning mounting height. design temperatures. Do not install units where combustible vapors are present. Select type of control suitable for the heaters and the specific application. 2. clearance to Combustible materials and general layout of the building. that the designer or engineer must follow the manufacturers design and layout suggestions in order to be protected by their guarantee. 5. Calculate building transmission losses using ASHRAE design procedures in Chapter 25 of the 1985 ASHRAE Handbook of Fundamentals [1]. 7. 6. 4. Calculate total heat loss by adding together the transmission losses and infiltration and ventilation losses. 1. Compute the air infiltration and any forced ventilation loads using procedures given in Chapter 22 of Ref. Take into account the manufacturer's recommendations and requirements. The following basic steps are suggested as a design procedure. This should take into account the mounting height.A design procedure is presented below for gas and electric infrared heaters. 3. Determine the number of heaters by dividing the total load from (4) by the heater size selected in (5). Reduce this number based on the information given in Figure 28 and the estimated infiltration air changes per hour. reflector design (they should be designed and placed so that no direct -142- . Determine if the building and/or operations are suitable for infrared heaters. reflector style. 1. Select heater size or sizes and type of control. It must be stated. Determine heater placement using the manufacturers suggestions regarding mounting height. however. distance from combustibles. Select and locate thermostats to control zone loads and provide uniform heating. Avoid having heaters too close to structural members. brackets.) and heaters in the interior of the building are usually mounted horizontal with appropriate reflectors. They should be mounted according to the manufacturers recommendations. 8. Comply with all of the manufacturers installation and operation instructions. In unvented gas fired systems check for condensation possibilities. and not in direct contact with cold outside walls. they should be about 5 feet from the floor. flexible gas lines and flexible electrical conduit. Use manufacturers recommendations concerning mounting devices. and light fixtures. 10. 11. storage racks. Mclntyre [23 and 24] used a theoretical -143- . Determine the method of mounting heaters using manufacturers recommendations. 9.infrared radiation falls on the walls and that the floor is covered with direct infrared radiant energy in proportion to the building heat loss). fork truck travel. Generally. They should avoid interior obstructions such as cranes sprinkler systems. 6. Always conform to local codes. Check design values of inside surfaces and roof with the deWpoint temperature of the inside air. Perimeter mounted heaters are usually angled toward the interior of the building (at about 30 deg.5 . out of direct view of the heaters. Determine minimum air needed to dilute CO2 in unvented units to a safe level. and building dimensions.Other Design Procedures A few other suggestions were found in the literature for modification of the design heat load for a radiant heating system when the design heat load is calculated for a forced air system. In another unpublished discussion of this problem. floors or ceiling.air systems. For about the greatest change illustrated. (2) theoretical studies showed little difference (5%) in the power required to maintain comfortable conditions in residential size and types of rooms with either radiant or warm air heating. -144- . he also looked at estimated energy requirements for the various types of systems. a combined modifier of 0. He came to the following conclusions: (1) the simplified CIBS Guide method and the computer model show very good agreement. This is somewhat higher than what others have calculated. this model is considerably different than what others propose.computer program and a simplified method in the CIBS Guide to compare the difference in design heating load for radiant and warm. These are all based on theoretical calculations and show the same trend of a decrease in the design heat loss as infiltration increases and an increase in the design heat loss as more radiant energy falls on the walls. In another paper by Harrison [25]. These contain multiplying factors to be applied to transmission losses and air change losses. This is comparable with the value of a 17% reduction given in" Figure 28. however. (3) radiant heating was more economical (5-20%) than forced air heating systems in large spaces with high infiltration rates.82 was given. the author calculates a percent reduction in the design heat loss of about 12% at one air change per hour. In addition. the differences between design heating loads for convective and radiant systems in discussed. and trade associations. measurement instrumentation and miscellaneous items. Approximately 75 positive responses have been received from these inquiries.SUMMARY OF MANUFACTURERS SURVEY Over 320 letters have been sent to the following groups: members and corresponding members of TC 6.5. The majority of the information has been specific details about products and some information on design procedures. The remainder of these responses discussed various aspects of the radiant heating field such as controls. This information has been reviewed and has been incorporated into this report. laboratories. consultants.0 . Approximately 15 to 20 responses were for medium and high temperature electric and infrared units. -145- . equipment and instrument manufacturers or suppliers as well as installers of radiant heating/cooling equipment.7. About 10-15 responses concerned radiant ceiling/floor panels. materials of construction. The dynamics of the system are important during transient load situations in order to estimate comfort conditions and energy requirements. The heated concrete floor and embedded heaters in plaster ceilings were found to present the slowest response times and require more sophisticated control systems to account for temperature lag.SYSTEM DYNAMICS System dynamics for heating and cooling enter into the calculations only when actual operation is considered and not when design heating loads are being calculated. The hydronic metal ceiling panels and high temperature ( infrared) systems do not appear to present any dynamic problems if properly designed and controlled. This is discussed in several references [19.26.8.0 .28].27. -146- .21. 9.2 . 9.0 . There appears to be a lack of information on what effect heating and cooling surfaces have on air temperature stratification. but no general calculation criteria appear to be available. 9. Most manufacturers do not present this data and some estimate it at between 0. Four of these areas enter directly into the calculations which have been made in this investigation.9. There do not appear to be reliable data for surface emissivities of radiant heating and cooling surfaces.RESEARCH NEEDS There have been six areas which were identified as requiring additional research information.95. 9.Comfort During Radiant Temperature Asymmetry. When surfaces (panel heating units or glass surfaces) at different temperatures than a surrounding surface in the same plane are present. Further research needs to be done in these areas. It is estimated that this unknown effect could influence the results by 5 to 10%.Air Temperature Stratification. reliable data does not seem to be available.3 . long thin heating surfaces). Some measurements have been made [18] for specific situations . There is insufficient information available on the effect of geometrical considerations for convection coefficients from surfaces at temperatures different than the air temperature.14] on the effects on comfort due to radiant temperature asymmetry . It would be useful to have design charts giving values of radiant temperature -147- .87 and 0.Convection Coefficients. There is some data available [7.Surface Emissivities.4 . There is also a lack of reliable data for the size effect of surfaces exchanging heat by convection with the air (tall walls.1 . This value has a strong effect on the area of heaters required for radiant systems. but additional information is needed. Heated Floor System With the reported discrepancies in the ASHRAE design procedure [20. Some applied types of heat transfer analyses involving the interac- tions between convection and radiant exchanges at surfaces with the material having a heat capacity are needed.5 .21. Some additional work is needed in this area. -148- .22] additional work should be done to correct the estimations of upward heat floWj downward heat flow.6 . 9. Very little information is available on radiant heating system dynamics and how this affects the energy requirements for radiant heating and cooling systems. 9.Radiant System Dynamics.asymmetry for specific geometries or giving limiting values of heat flux (Btu/hr ft^) for comfort conditions. and edge heat losses for different geometries and tube spacings. 1972.10.of Heating. Atlanta. V. ASHRAE Trans. 1984. ASHRAE Trans. 1. 1. and Nielsen. II. 0. R. Jr. G. 1968. H. F.REFERENCES ASHRAE.. H. American Society of Heating. 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Piping and Air Conditioning. M. Heating Piping and Air Conditioning. IHVE. New York. Merl. 1977. "Losses from a Floor-Type Panel Heating System. L. Hunter. July 1951. .. "Aluminum Ceiling Panels for Heating and Cooling. Inc. New York. D. June. A Graphical Design Procedure for Radiant Panel Heating.E. June 1951.. H. Hutchinson. C M .. Design of Heating and Ventilating Systems. Nottage. Franks. "Optimum Panel Surface Distribution Determined from Human Shape Factors.W. F. Piping.. L.. W. Piping and Air Conditioning.C. C. W. L. F. F. . L.. Piping & Air Conditioning. Hutchinson. ASHVE. Huddleston. E. Piping and Air Conditioning.. Irwin. 1948. F.. IHVE Guide. Huebscher. R. F. Humphreys.W. "Thermal and Other Properties of Building Structures". and Locklin.S." Heating. Piping & Air Conditioning. L. F. 1947.. "Laboratory Studies of the Thermal Characteristics of Plaster Panels. Trans. Hulbert. Mills. "Controllable and Efficient Infrared Radiant Heating". A3. V. and Baker.New Infrared Heating System Solves Old Problems".F. and La Tart. L. Hutchinson." Heating. F. A. de chauffage par rayonnement a' basse > "Le Procede Calendal: temperature". 82." Heating... Vol. A-ll . L. Piping..Janssen.. 1948. Piping and Air Conditioning. and Schreiber..... "Radiation Cooling of Structures with Infrared Transparent Wind Screens". May. 29... Sebal. Pg. Vol. A. Macey. Institute of Fuel Journal.. "Thermal Comfort and Energy Consumption in Winter Conditions -. 1949. July. No. R." ASHRAE Trans. Refrigerating Engineering. G. Vol. & Air Conditioning. 438. W S Kongres . T. Madsen.W S Messe.. T. pg. "Radiation Characteristics of Gas Infrared" Journal of Engineering Physics.. Korsgaard.. G. MacLeod.. -.Continuation of the Experimental Study" ASHRAE Trans. ACEEE 1984 Summer Study on Energy Efficiency in Buildings.Mortensen. "Field Evaluation of High Temperature Infrared Space Heating Systems. & Air Conditioning. 1975. J. July 1947. Korsgaard. 86. J. 1. 1985. "Necessity of Using a Directional Mean Radiant Temperature to Describe the Thermal Conditions in Rooms". February 1950.." New and Existing Single Family Residences.. V. Gunnarsen." Heating. p. Consultants Bureau. "The Mechanism of Heat Transfer Panel Cooling Heat Storage Part II. C. Piping. Pt. 1395-1400. "Criteria for Mechanical Energy Saving Retrofit Options for Single Family Residences. John E. V. Leopold. 1951. UCLA. Lebrun. "A New Radiometer Measuring Directional Mean Radiant Temperatures". June. January. 1949. Heating." Clima 2000 . "Design Factors in Panel and Air Cooling Systems. 1980. N. Khudenko. "Comfort Limits During Infrared Radiant Heating of Industrial Spaces. Refrigerating Engineering.. Johnson. Pt. V. August. Reuve de 1'aluminium.E. 1949.. 2." Heating. S. 5. Dominique J. F. March 1975. C. "Heat Loss Through a Solid Floor". Heating Pining and Air Conditioning. 1979. Piping and Air Conditioning. ASHRAE Trans. and Marret. Proceedings of the Second Workshop on the Use of Solar Energy for the Cooling of Buildings. and Eves. 369. C. E. Vol.No. Leopold. "Baseboard Radiation Performance in Occupied Dwellings. S. 1976. 4. 1. Charles S. Solar Radiation. Pt. June. Langkilde.. L. H. Lorenzi.Indoor Climate. "The Mechanism of Heat Transfer Panel Cooling Heat Storage". Vol 85. H.. 141-1565. S. Kweller Esher. "Performance of an Electrical System of Panel Heating with Four Stages of Insulation. "Definition and Measurement of Local Thermal Discomfort Parameters".. Leopold. Jean J. 22-128. D. The Building Services Engineer. 1974. "The Efficiency of Radiant Heat Sources". Capenhurst. A. Batiste Publ. . London. A. 81. ASHRAE Trans. Mclntyre. V. "The Relative Effects of Convection and Radiation Heat Transfer on Thermal Comfort (Thermal Neutrality) for Sedentary and Active Human Subjects". Illuminating Engineering Society.. pp. J. 1976.. D.. Vol. J. A. 8. "Warm Air and Radiant Heating: Steady State Power Requirements". No. Mclntyre. 1. England. October.A. Mclntyre. (NTIS-PB83-231506). March. 226-34. A. "Evaluation of Thermal Discomfort". Feb. 2. 1985. 1976. 3. . January-February 1984. Jr. 287-296. . R. P. (NTIS-PB 277-428).. T. England. Capenhurst. V. E. "Thermal and Comfort Sensations of Sedentary Persons Exposed to Asymmetric Radiant Fields". Mclntyre. 1984. D. A. C .. Vol. D. Electricity Council Research Centre. Building Science. The Building Services Engineer. Mclntrye. K. 1977 (NTIS-PB 277 115). D. 40. McNall. 74. 67-70. D. The Building Services Engineer. L. D... Mclntyre. 1. 1976." ASHRAE Trans. "A Manufacturer's View of Radiant Heater Control. Madsen. Pt. T. L. 1977. (NTIS-PB 85-189975). "Thermal Radiation from Lighting Installations". May. 1970. 41. IB. D. P. Ergonomics. Electricity Council Research Centre. 17. A-12 . Capenhurst. Pt. pp.. "The Thermal Radiation Field". "Warm Air Or Radiant Heating?" Building Research and Practice Vol. 91.A. 1968. 1973. 48-51. Ltd. D e c .. V. D. "Overhead Radiation and Comfort". . and Biddison. B. Vol. ASHRAE Trans. . Pt. "Comparison Between Operative and Equivalent Temperature Under Typical Indoor Conditions".. England. Electricity Council Research Centre. "Sensitivity and Discomfort Associated with Overhead Thermal Radiation". "Eight Hour Floor Warming: A Feasibility Study". Pg. Capenhurst. E. and Brailsford. "Thermal Comfort Measurements". Mclntyre.. A. R. Pg. No. No. A. and Schlegel. "Overhead Radiation and Comfort". Mclntyre. Olesen. McNall. England. April. E. D. Preston E. Mclntyre.. V.. W.. ASHRAE Trans. D. 1973. 20. Pt.. A. 9. 1. N. Mclntyre. Part I. Electricity Council Research Centre. Mclntrye. Kristensen. Jr. A. January 1977.. ASHRAE Trans.. V.. V.Madsen. Centre Scientifique et technique du bAatiment. 76. 3. 1973. . 82. Vol. 44.. "Radiant Heat from Lights and Its Effect on Thermal Comfort".. McNall. Vol. 1980. Moisan. G.. and Feyerherm.. Radiator Modelling. and Hwang. 1957. A. .W S Meese. Nevins.1. July-August. Mills. ASHRAE Trans. Stardard Publication No. Piping & Air Conditioning. p. June 1968.J. 1984. 1966. and Feyerherm. 1985. . A.A.. Refrigerating Engineering. C. "Comfort in Damp Cold Air with Radiant Spot Heating". Pt. 1950. "Heating of Industrial Buildings with the Help of Suspended Radiating Panels. CIB/RILEM Symposium on Moisture Problems in Buildings. "An Economic Study of an Electric Infrared Space Heating Installation." Refrigerating Engineering. HE 3-1983..Part I. L.. Rotterdam.Heating. L. 1974.A. Heating Piping and Air Conditioning. ANSI Z 223. March 1958. G. 39-49. National Fuel Gas Code." Clima 200 . Clarence A. . "The Effect of Floor Surface Temperature on Comfort Part III. M. Nakanishi. Cleveland. Mills. November 1950. A. 2. Schutrum. "Effect of Floor Surface Temperature on Comfort Part IV: Cold Floors".. C ." Clima 2000 Heating.F. ASHRAE Trans. R. A.V.. W S Kongres ..D. 1955. Building Research. . J. G. 149-153. Min. Oct. Mills. p. 1973. E. 1964.. and Vouris. 1985. and Flinner. Infrared Application Manual. Fan. V.A. C. Mills. NEMA.. . 1966. 6. 0. R. "Sensible vs Latent Heat Removal in Radiant Cooling. "Year-Round Residential Conditioning By Reflective Radiation. Proceedings of. J. Jan. A. American Gas Association. Ventilating and Air-Conditioning System. B. Vol. Solovyov. Shilkrot. V. No.. Parmelee. Clarence A. . T. May 1956. Vol. M... 11. November. M. Ventilating and Air-Conditionign Systems. 6.. R. A. 1984." ASHRAE Journal. Building Science. M. A-13 . E. "Criteria for Thermal Comfort". College Age Females". 73. Naumov. G. A. K. National Electrical Manufacturers Association." Refrigerating Engineering.. Nevins. L. T." Refrigerating Engineering.. Washington. 70. Nevins..O. C. 58. "Residential Cooling by Reflective Radiation. Pereira. DC.. "Natural Convection and Radiation in a Panel-Heated Room.Michaels. J. "Simulation of a Hydronic Heating System. Ohio. L. 8. "Simultaneous Control of Temperature and Humidity in a Confined Space . Vol. . "Reflective Radiant Conditioning Can Provide More Comfort at Less Cost". Morant. and Strengnart. Vol. R. N..C. "Effect of Heated-Floor Temperatures on Comfort". G. Kongres W S Messe. W S . Nevins. Morelli." Heating. and Lebrun. 1983. the Second Int. Nevins.R. Franks. 91. Perry. 89. L. . R. Feb. 1980. IA-19. V. T. Teledyne Laars. B. 2B. "The Effect of Floor Surface Temperature on Comfort Part 1. 1985. Vol. W. Nottage. B. Vol. V. Pedersen. on Indust. . Pt. College Age Males". Pt. V. V.V. P. Business News Publishing Company. Bunkofske. Fred. "Space Heating Dynamics". A-14 . L. Pt.I. and Kesselring.. "Experimental Study of Radiation and Convection Heat Exchange in Rooms for Energy Analysis Program Models. Olesen. Pam.. MI. W S Kongres .. ASHRAE Trans. January 1984. Pierce.. "Application of Fin Tube Radiation. Vol. Hulbert. Burner Survey for a High Efficiency Gas-Fired Heating Unit. V. 1... J. G. "Forced Convection Heat Transfer from Flat Surfaces".. . Alzeta Report No. 1983. 93. 86. B. J. 1. Plattls. V. and Huebscher. E. K.F. G. C... 1. "Causes and Prevention of Air Temperature Stratification". Olivieri. Schutrum. 55. Prince. ASHRAE Trans. F." ASHRAE Trans. May 1953.. "Heat Flow Analysis in Panel Heating or Cooling Sections. B." Clima 2000 . E. and Scesa. 1963.." Heating. Feb. . 1963.. .. Sept/Oct. Birmingham.. .. October 1962. Olivieri. T. Thorshauge. 1983. April.B. 1971.. Parmelee. C.J. 1964. Pfafflin. 70. ASHRAE Trans.. Cunningham.H. Peach. IEEE Trans. R. J. G. 5. Vol. 1947. J.. 2. May 1982. ASHRAE Trans. Pt. 87. and Singh. J.. How to Design Heating-Cooling Comfort Ceilings. J. 1987. V.. Pg...J.. J. W. Olesen." ASHRAE Journal.D. J. Leverenz. "Thermal Comfort in a Room Heated by Different Methods". M. D. Piping and Air Conditioning. Michaels. and Nielsen.E. "A Simplified Calculation Method for Checking the Indoor Thermal Climate"...W S Messe. H. W. 1985. V. Canadian Builder. B. 39. 1981. J.E. ASHRAE Trans. Plant Engineering. R. "Radiators and other Convectors". R. A. . Appls.. "An Analysis of Heat Losses through Residential Floor Slabs.. "ASHRAE Journal. "Selection and Application of Overhead Gas-Fired Infrared Heating Devices. Prince. 2. D. S. J. "A Computer Program for Radiant Cooling of High Bay Buildings". E. Spitler. R. R. 84-706-104. Pt. 1972. and Berg-Munch. No. ASHRAE Trans... "Radiant Spot Cooling of Hot Working Places". and Feyerherm. G..D. B. "Where Polyethylene Pipe Challenges Metal for slab Radiant Heating". Mortensen. Olesen.. B. H. L. J.Building Design and Performance. 53. Rapp. 1. Borje E. 1. 93. 25.. 14-18. July. B. W. W. November. 2. Uppsala. 1944. Fred J. July-August 1984. & Air Conditioning. Sartain. A. J. ASHVE Trans. Heating and Refrigeration News. F. 86. October. t Radiant Floor Heating. E. W. 1955. "Configuration Factors for Radiant Energy Interchange with Triangular Areas. P... Sweden. S.. W. No. John Wiley & Sons. NY. Sanford. 73-83. 2. & Air Conditioning. Vol. H." ASHRAE Trans. W. Hans. and Hutchinson. Sauer. Sartain. Rapp. and Andrews. Institution of Electrical Engineering Proc. Sartain.. Vol.. . Dec. Pg. 1974. . and Harris. Piping. Pg. Len. Vol. "Performance of Covered Hot Water Floor Panels . Piping." ASHRAE Trans. 1967." ASHRAE Trans." Heating. January 1954. Rohles. and Harris. 1966.. Pt. Heating.. F. March 1985..Thermal Characteristics"... "Configuration Factors and Comfort Design in Radiant Beam Heating of Man by High Temperature Infrared Sources. Janca Enterprises Ltd. . 1947. 1968. 1956. "The Effect of Wall Reflectivity on the Thermal Performance of Radiant Heating Panels". and Lofstedt.. 1957. "Heat Flow Characteristics of Hot Water Floor Panels. 72. September. "No Problem with Radiation" The Heating and Air Conditioning Journal.. and Hutchinson. Piping. 1980. Pt. Saunders. 18.... ASHRAE Trans..S. E." ASHRAE Journal.Part II . Roots. 73. 114. "Selected Segment Hydronic Heating System. B. H. "Analysis of Free convection and Radiation Heat Transfer in Valance Heat Exchangers. "Radiant Cooling Panel will get Tryout in U. No. Panel Heating and Cooling Analysis. F. George and Gagge. Pt. Harry J. 7. and Harris. & Air Conditioning." Heating. 80. F. "Infrared Heating for Overall Comfort. "Radiant Drafts from Cold Ceilings. S.. S. George. 21. L.. "Electric Space Heating with Active Boundary Members". Pt. Piping..f "Performance of Covered Hot Water Floor Panels .. "Optimum Surface Distribution in Panel Heating and Cooling Systems". Vol.. 1. Pt. 630. 1985. 54. F. A-15 . Ronge. Raber.Prince. Plasco Manufacturing Ltd. "Temperature or Temperament: A Psychologist Looks at Thermal Comfort".. 1. Troup Publ. Vol. & Air Conditioning. L. James D. W. Raber. 1985. J." Heating. .Part I . Jr. Air Conditioning. 1987.. E." Energy Conversion Mgmt. No. Vol.Room Conditions. ASHRAE Trans. Rickman. L. W. V.K. E. October 21.". 1967. V. N. 1957. . and Vouris.E. Schutrum. . and Min.. T. 77004.. C ." Heating.F. 1985.. L." Heating. 2." Heating. Schutrum. 1968. and Min. T C .. T." Heating. (UNPUBLISHED) Simmons. 1968. E.. & Air Conditioning. . J.By a Cooled Ceiling Panel. P. H.. "Lighting and Cooled Air Effects on Panel Cooling. December. "Heat Exchanges in a Floor Panel Heated Room. J. B. "The Effect of Asymmetric Radiation on the Thermal and Comfort Sensations of Sedentary Subjects". J. Jr. . Cleveland. E.G. Parmelee." Heating. Schlegel.. August.Schlegel. and Olivieri. F. August. R.F. John. 1981. P. A-16 . L. University of Houston. ASHRAE. 2.F. Schutrum. January. Schutrum. 1954. ASHRAE Trans. C M . F.F. L. Ohio. "Effects of Non-Uniformity and Furnishings on Panel Heating Performance. Vol. Piping and Air Conditioning. Parmelee.. February. Piping and Air Conditioning. Cleveland. C. June 1968. November. C and McNall. Cleveland. V. 88.. L. Piping & Air Conditioning.. July 1953. 1956.Panel-Heated Room. M. "The Effect of Asymmetric Radiation on the Thermal and Comfort Sensations of Sedentary Subjects". & Air Conditioning. C. II. Final Report. Department of Mechanical Engineering. T. B." Heating. and Min. "Further Studies of the Thermal Characteristics of Plaster Panels. 1955. Pt. Ohio.." ASHRAE Journal. C . "Performance of Polybutylene Pipe in Concrete Heating Panels". V. and Olivieri.. Ohio. Piping. V. Singh. Schutrum. C . Piping and Air Conditioning. F. T.F. Singh. "Preliminary Studies of Heat Removal. Piping. Vouris. J. 1982. Piping and Air Conditioning. L. Humphreys. and Tezduyar. Schutrum. and Humphreys. Piping and Air Conditioning. Lienhard. 74." Heating. G. "Cold Wall Effects in a Ceiling . ASHRAE Trans. CM. C M . "Five Years Operation of an Industrial Infrared Heating System. and McNall. Humphreys. 74. Report No. Texas. Shamsundar. T. "Effect of Radiant Cooling Panels on Temperature Stratification under RP-260". "Heat Exchanges in a Ceiling Panel Heated Room. Schutrum. "Effect of Radiant Cooling Panels on Temperature Stratification" ASHRAE Trans. June 1953. J. "Effects of Room Size and Non-Uniformity of Panel Temperature on Panel Performance. D. Houston. .. L. ." Heating. and Humphreys.. V. Schutrum. L. 1952. 1954. J. L. . September.. Pt.. A. 2. Trewin.. E. "Thermal Design of Warm Water Ceiling Panels. F. Staff members of the ASHRAE Research Laboratory. "Radiation Heat Transfer at a Surface Having Both Spacular and Diffuse Reflectance Components".. 1986. and Parmelee. J. Vol. Heizungstechnik Kg. M. "The Effect of Floor Surface Temperature on Comfort: Part III. A. Industr. 12.L. pp. A. Taylor. M. Vol. M. D. 2. Int. Vol. 1980. D. 93. Herriot. V. 1. "Spot Cooling System Design". Stevens. 1977. Pt. Spangler.. 58. M. 63. Vol. 1987. V. Jnl. S..... Pt. Subcommittee of TAC.. W... Nelson. R. A.. R. 1955. C. . ASHRAE Trans. 1969. Efficient Computation of Zone Loads. Stevens.. M. & Air Conditioning. 1965. V. 1. N... "The ASHVE Environmental Laboratory".Industrial Radiation Thermometry". 149-165. Oct. Heating and Ventilating. Spolek. K. ASHRAE. December. "Industrial Climate Control Versus Radiant Heat". G. Planungsunterlage Fur Ingenieure. ThermoLutz GMBH and Co. ASHRAE Trans. 72. . G.. B. T. V. and Rae. "Subjective Warmth in Relation to the Density. E.. March. of Heat & Mass Trans. R. F. 143-146. 1980. Mechanical Engineering. "An Experimental Study of a Multipurpose Commercial Building with Three Different Heating Systems"." ASHRAE Trans. Tenney. Air Conditioning. H. and Walton. P. Duration. and Marks. E... M. A. L.Smith. W. M." ASHRAE Trans. 1965. 92. 86. 8. The Elderly". 76.. Piping. Sparrow. Humphreys. 1986. Heating. V. "Radiant Space Heating". Langdon. 22. ASHRAE Jnl. "Red Hot and Hotter . M. Nevins. Pt. 1966. G. Tasker. pp. and Pate. 1957. Joseph C . 24.. B. C. Pt. J. May/June. Environmental Research. "Airflow in Rooms with Baseboard Heat: Flow Visualization Studies". and Michaels. S. 1. Tredre. Trans. Jan. V. R. Pt. 1. and Areal Extent of Infrared Irradiation.. A.. . Jan. t Thermo Lutz. B. "The Quantitative Assessment of Thermal Comfort". E.. . C. Springer. Y... Vol. ASHRAE Trans. E. Med. 1970. 1965. V. G. 1973. "Assessment of Mean Radiant Temperature in Indoor Environments". "Thermal Design of Warm Water Concrete Floor Panels" ASHRAE Research Laboratory... Lawrence. Marks. H. M. Building and Environment.. A-17 . and Gagge. Feyerherm. and Al-Hukail. Brit. 1952. M. Building Materials. "Patient Comfort and Radiant Ceiling Heating in a Hospital Ward". Sofrata. Ill... Subcommittee of TAC. Piping and Air Conditioning. and Lin. and Low." Heating. R. F. Sowell. 1986. "A Numerical Study of Heat Transfer in a Hydronic Radiant Ceiling Panel"." ASHRAE Trans. V. George N. 91. Weida. New York. 1985. ASHRAE Trans. 1. "An Experimental Study of the Transient Response of a Radiant Panel Ceiling and Enclosure". Vol.. Pate. 1986." ASHRAE Trans. 62. A-18 . M." ASHRAE Journal November. and Peterson. Walton. "An Experimental Study of an Installed High Temperature Radiant Heater and Enclosure. C. Z. P. Z. and Pate. Vol. California.. R. G. R.. Pate M. A. "Heating a Basementless House with Radiant Baseboard. Zawacki.E. ASHRAE Trans. B. B. Zhang. 1.. Anaheim. 2... "Life-Cycle Cost Analysis of Hydronic Radiant Panel. 1. Z. D. H. 92. R. Proceeding of the 1986 ASME Solar Energy Conference. Pate.. "A New Algorithm for Radiant Interchange in Room Loads Calculations". Weigel. "Development of a Standard Test Method for Measurement of the Radiant Heat Output of Gas-Fired Infrared Heaters". 86. 1985.Trewin. 1949. Heating and Refrigeration News. R. 44. N.E... Pt. "Analysis of Slab-Heated Buildings.. V." ASHRAE Trans. Zhang. 1938.. 93. HTD-Vol. J.. SED-Vol. 21. 2. Heat Transfer and Fluid Flow in Solar Thermal Systems. ASHRAE. 1. R. Zhang. M. "Radiation and Convection from Surfaces in Various Positions". B.. and Macriss...M.. "Control of High Intensity Infrared Heating. and Pate. Oct. ASHRAE Trans. V. . C.. Z... Pt. "Underfloor Radiant System Uses 86H Supply Water". "A Performance Evaluation of a Residental Solar Hydronic Radiant Heating System". 1985. S. S.. 1986. .. 2. Walker. W. V.A. Pt. and Shapiro. M. Numerical Methods in Heat Transfer. No. T. H. Air Conditioning. Huang. Pt. B. 1980. and Nelson. 55. VanGerpen. Vol. New York. T. H. Wilkes.... April 14-17. Zhang. Pt. 1962. Pt. "An Experimental Study of a Residential Solar System Coupled to a Radiant Panel Ceiling". ASME. 1987. and Harris... 1." Trans. 92. and Nelson. 1986. M. 1986. ASME. B.F. Vol. Liu. V. ASHRAE Trans. 92... APPENDIX B ANNOTATED BIBLIOGRAPHY " < ! • ' la B-l m . a short discussion concerning each article has been prepared. In many instances. this is a reproduction of the abstract or conclusions of the article. A) Load Analysis and Modeling B) Convection Coefficients C) General D) Comfort Conditions E) Thermal Comfort . The categories in this Annotated Bibliography are as follows.Radiant F) Floor Panels G) Panel Heating and Cooling H) Infrared Heating I) Design Procedures J) Energy Consumption K) Transient Effects L) Instruments M) Controls N) Spot Heating and Cooling It should also be noted that there are some secondary references available from most of the entries in the Bibliography by examining the references for that specific entry. but are obtainable by locating the article given in the Bibliography and then locating the references in that specific article. B-2 . All of these secondary references are not listed.ANNOTATED BIBLIOGRAPHY For all of the entries given in the BIBLIOGRAPHY. V.ANNOTATED BIBLIOGRAPHY A) Load Analysis and Modeling 1. Pt. Also. 1978. 84. It contains algorithms for calculating the following items: solar load on walls. The modeling procedures and the transient comparisons are described in detail. This study has produced a numerical solution to the process of transient thermal stratification occurring in factories." ASHRAE Trans. roofs and glass.Mathematical Mode". D. A computer program based on this model was used to generate predictions of cooling loads created by lighting for a variety of building-lighting arrangements. "Thermal Stratification in Factories . psychrometric property calculations. This indicates that the technique of load reduction by stratification can work whenever an undisturbed strata can be formed. pt. ASHRAE. H. external shadow calculations. Procedure for Determining Heating and Cooling Loads for Computerizing Energy Calculations . exhausting needed ventilation for the cooled space through the roof can B-3 . radiation shape factors. Vol. Atlanta. airflow paths and rates. and Gorton. This ASHRAE publication presents algorithms for calculating the heating and cooling loads for structures. 2A. A mathematical model describing the spatial and time distribution of energy produced by lighting has been constructed. Robert L. . and other miscellaneous algorithms. and lighting duration. 1975. Comparisons of model predictions are made with published experimental steady-state and transient results. and Green.. convection heat transfer coefficients. * 3. The results indicate that when varied over a reasonable range.Cooling Loads and Temperature Profiles. American Society of Heating... "The Impact of Lighting Fixtures on Heating and Cooling Loads . physical model with thermal plumes rising from the light fixtures in a stratified layer between the lights and the ceiling. The model is based on well-known heat-exchange computation procedures and contains no empirical coefficients from lighting energy-transfer experiments. ceiling height is of no importance in the system in that it has no influence on cooling load magnitude. 1983. 1. Refrigerating and Air Conditioning Engineers. ASHRAE Trans.Algorithms for Building Heat Transfer Subroutines. This model was used to generate values of room and total loads and to present these in a form or forms suitable for use in design. The analysis was restricted by considering the light fixtures as defining the interface between the conditioned area and the stratified area. GA. The solution was based on a simplified.. * * Ball. D. * * Beier. Richard A. * 2. 89. Comparison of results from the computer model to measured values from one factory show acceptably close agreement. January 1980. This is a German article describing the radiant energy balance in a room which is receiving solar radiation through the glass and on its exterior surface. they are thought adequate for nearly all engineering work..' "Configuration Factors for Radiant Heat Transfer Calculations Involving People". 1981. R. For greatest accuracy. If more precise information is needed for a specific individual. San Diego. * 6. which acts as a clearinghouse for all radiative exchanges. While people of unusually stout or slim build may deviate somewhat from the empirical equations. 39. V. The "MRT View Factor" method presented in this paper couples each surface in a room to an MRT node. An "MRT Method" of Computing Radiant Energy Exchange in Rooms. Jnl. W.. * * Dunkle. * * Carroll. Feb. August. B. Joseph A. CA. pg. The adjustments also happen to improve the accuracy of the conventional view factors implicitly assigned by MRT methods. Errors do occur in the "implicit view factors". 4. and these errors can be compensated for if necessary. but errors such as this are inherent in any method which overlooks the gory details of the enclosure geometry. * 4.05-.. No. 203-210. An upward adjustment in the coupling between each surface and the MRT exactly cancels that surface's self-weight in the MRT. Vol. * 5. recourse can be had to experimental measurements. Example temperature profiles are given in the article. and Mommertz. This method is inherently free from heat balance errors and errors in the overall radiative coupling of each surface to its environment. "Untersuchung von Flachenheiz Systemen Mit Der Thermoelektrischen Analogie". It uses the electrical analogy for surface radiant exchanges. Systems Simulation and Economic Analysis Conference Proceedings. The effects of surface emittance and air emittance (typically . More extensive field comparisons need to be made and detailed laboratory investigations of some of the basic assumptions used in model development are required before the model can be applied with full confidence. Elektrowaerme. * * Endreb. Six configuration figures are presented for a "standard person". B-4 . radiation coefficients can be varied with temperature. 19 6 3.15 in residences) are modelled without difficulty. Coplanar surfaces cause the largest errors. Von H.reduce the cooling load. It is thought that the point and area configuration factors for people reported in this paper will prove very useful in heat transfer problems involving people. of Heat Trans fer.. ASME Trans... Results from a number of situations of interest were presented. In the calculations. 1983.. July 1958. "Deriving Radiation View Factors within a Triangular Cross-Section Residential Attic". 1A. to account for storage effects in the structure and in the space air.& Air ConditioninE. The same basic technique is used in all calculations with only minor alterations. A computer program is available to solve radiation view factor equations for some representative attic configurations. the numerical operations are lengthy.7. Since the major inconvenience in the use of the method is in connection with the calculation of absorption factors. T. This is the standard practice for the similar quantity. 90. and Leard. 1984. cross-section attic space is derived. V. The analysis is direct and consistent because of the way in which basic quantities are defined and is similar to the approach used for black surfaces. A simple set of nodal radiation heat transfer equations is also included to demonstrate the application of view factors to an attic floor and two attic roof sections radiating to one another. and to determine the resulting space air temperature profiles and space-cooling loads. These view factors allow nodal radiant energy-transfer equations to be solved for radiant heat exchanges between the attic floor. A description of a computer program capable of computing time-varying cooling loads for a stratified-cooled space has been presented... * * Gorton. pt. R. A simplified analysis of radiant and convective heat transfers within an idealized attic space shows radiative heat transfers to be on the same order of magnitude as convective heat transfers.. IB. The involved calculations arise in the solution of the equations for the absorption factors. New York. * 8. ASHRAE Trans. angle factor. roof sections. ASHRAE. Benjamin. Vol. symmetric. A. For a many surface enclosure. endwalls. as in any calculations which account for many processes simultaneously. L. "A Computer Program for Air Temperature and Cooling Load Determination for Stratified-Cooled Industrial Buildings". The program is structured to accept hourly changes in internal loads and environmental conditions. Ford. J. A new method of thermal radiation analysis for non-black surfaces is demonstrated by calculating various radiant energy transfer rates in a room. A complete set of radiation view factors within a triangular. * * Gebhart. Pt. Heating Piping. K. "A New Method for Calculating Radiant Exchanges". 89. Trans. and soffits. Ithaca. the question arises as to whether or not the values of the absorption factors could be presented graphically. * 9. . this is necessarily a machine calculation. B-5 . L. Gorton. There is an accompanying reduction in lower-zone cooling load. and Sassi. 2. Air-Conditioned System: Part 1-Model Studies". M.. Gorton. The results are limited to conditions of uniform floor-level loads (no thermal plumes) with "perfect" air distribution (uniform diffusion of air in the cooled space and no interaction with the upper-zone stratification pattern). and thermal variables. Loads originating at the roof reach the cooled space by radiation to the lower zone. R. "Determination of Temperature Loads in a Thermally Stratified. flow. 1982. The major conclusions reached from a qualitative evaluation of the test results are: 1. ASHRAE Trans. 88. . A computer program for use in determining space air temperatures in a stratified-cooled scale model space is described. 3. 1982. The level of temperature is changed. "Determination of Temperature Profiles and Loads in a Thermally Stratified. Pt. 11. however. The supply air temperature and flow rate do not influence the temperature patterns in the space. the reduction is small and high-level exhaust cannot be considered as a significant means of load control. for a given load. Program results are compared to experimental results B-6 . Air Conditioned 2 .Program Description and Comparison of Computed Results". Experiments have been run in a model chamber to provide information on temperature patterns that develop in stratified-cooled systems as a function of various combinations of geometric.10. The space temperature profiles and cooling loads are relatively insensitive to the height of the lights and to their position relative to the cool zone and the roof. Convection from the roof is effectively blocked by the stagnant. M. . equation formulation. by changes in temperature and flow rate but not the general shape of the temperature profile. Program logic. and calculational models of the various heat transfer and flow processes are presented.. V. R. 5. and Profiles and System: Part and Measured Sassi. Loads originating at the lights reach the cooled space by radiation and by means of air circulation caused by free convection currents around the lights. Exhaust of air from high in the structure in the stratified zone provides a temperature reduction there proportional to the exhaust flow rate. 6. 2. M. high-temperature air layer adjacent to the roof. ASHRAE Trans. The space temperature profiles and cooling loads are relatively insensitive to the height of the roof and to its height relative to tKe cool zone. 4. M. 88r Pt. L. 2. V.. Pt. Feb. * * Harrison. comparisons within 1HC being obtained in most cases. 1974. each to serve different ends. and the other factor is related to the infiltration difference between air convection systems and radiant systems. The program is shown to be capable of producing acceptably accurate estimates of the measured temperatures. The other six are implemented on large computers. The Building Services Engineer. CA. T. All were considered in light of their completeness. Keith and Lydon. Many estimated air temperature profiles are given.. * 13. These comparisons are presented for a range of geometric and load variables. Vol. Jr. V. * 12. 19-23. "Computer Simulation of Radiant Heat Load and Control Alternatives". B-7 . Two of these are meant for calculators. May 1975.. E. To that end it has discerned eight basic algorithmic types. Batiste Publications Ltd. Robert L. Journal of American Industrial Hygiene Assoc. * * Harris. "Calculation of the Heat Requirements of Rooms". 35. A digital computer program calculating the exchange of radiant heat between an absorber and a multiplicity of radiators in its surroundings is described. The contribution of each individual radiator to the heat load on the absorber is obtained. Repeated computer runs with simple changes in input data may be used to simulate various control alternatives. One factor relates to the additional heat transfer through the wall due to radiant panels. R. The program permits mathematical assessment of radiant heat exchange problems for which manual calculations would be prohibitively time-consuming and costly.. 43. This paper has endeavored to explore the published state-of-the-art of passive solar load models. This article describes two factors which can be used to modify the standard design heating load when radiant heating systems are used. * 14..obtained in the model chamber. San Diego. pg. A matrix of absorber locations within a workspace having radiators in its surroundings may be specified. It has been found that all considered models make a trade-off between cost and model completeness. Comparison of Building Thermal Analysis "Methods.. permitting the construction of isopleths of heat load in the workspace for any elevation of interest. inputs and appropriateness of output to the various stages of building design. efficiency. * * Harrington. Systems Simulation and Economic Analysis Conference Proceedings. These would be useful in load analysis calculations. January 1980.. "Influence of Gaseous Radiation in Panel Heating".[E 0 (w/m 2 )/h 2 ] exp[-b x c ] and values for EQf b and c are given in the paper. Charles. Vol. * 16. 53. Lloyd M. W. "Necessity of Using a Directional Mean Radiant Temperature to Describe the Thermal Conditions in Rooms". Consultants Bureau. Piping and Air ConditioninE. 29. * * Hutchinson.15. F. 5. Very few details or algorithms are given. Vol. This expression i s : E(w/m2) .. 1972. ASHRAE Trans. only discussions on what must be done. Heating. Vol. A. * 18. * * Korsgaard. 2. gaseous radiation is. Because of cumulative absorption as associated with multiple reflections. A. This is a discussion on how to simulate a building HVAC System. Ft. vapor pressure (expressed in terms of room air temperature and of relative humidity) and surface temperatures. Equations are developed to permit evaluation of radiant exchange between certain systems of surfaces when they are reflective and are separated by an absorbing medium. Hedgepeth.. 1947. 1949. An a n a l y t i c expression i s obtained for the d i s t r i b u t i o n of irradiance over a f l a t object a t v a r i o u s d i s t a n c e s from the h e a t e r . responsible for reducing the effectiveness of reflective surfaces (when used as room surfacing in an attempt to reduce radiant body heat loss) to a negligible value. 1395-1400. June. 78. . V. Pg. * * Khudenko. Gaseous radiation does not appreciably affect either the panel size or panel rating for an ordinary panel heating system. "Radiation C h a r a c t e r i s t i c s of Gas Infrared" Journal of Engineering Physics.. ASHRAE Trans.. "A Thermodynamic Simulation of a Building Environmental Control System". but it does reduce the effectiveness of local (direct transfer) panels by about 10 percent. No.. The building did have a radiant ceiling heating system. A graphical solution is presented from which the equivalent coefficient for gaseous radiant exchange can be evaluated as a function of separating distance. The conclusions reached by this author are: B-8 . Results of an experimental study of the r a d i a t i o n c h a r a c t e r i s t i c s of commercial models of gas infrared heaters are presented. * 17. and Sepsy. but no information is presented. * * Mclntyre. Radiant heating was advantageous when only local warming of part of the room was required. * 20. The efficiency of the warm air system was reduced by temperature stratification if the inlet temperature was high and the warm air was inadequately mixed. causing an increased convective transfer coefficient. No. "Warm Air or Radiant Heating?" Building Research and Practice. Power requirements and heat loss estimates were determined using a simplified procedure together with a more detailed computer program. * 19. Capenhurst. When running comparative tests on different heating systems -especially when they include such types of heating systems where the thermal conditions in the room are regulated mainly by varying the temperature of larger or smaller parts of the inside surfaces of the room -.231506). for at any single point on a person's body the heat transfer conditions represent the combined influences of radiation exchange with those surroundings exposed to that part of the body.. but is a physical quantity that usually varies both with the place and the direction within a room. (NTIS . the possibility of savings would have to be set against high radiant asymmetries and an uneven temperature distribution over the room.PB83 . It was found that in a well insulated domestic room there was little to choose between the two types of heating systems. (2) Theoretical studies show little difference in the power required to maintain comfortable conditions in a domestic size room with either B-9 . In comparing warm air and radiant heating systems. England. The directional mean radiant temperature (DMRT) of a room with respect to a surface element. Electricity Council Research Centre. 1. A. Centre Scientifique et technique du bAatiment. Radiant heating was shown to be more economical than warm air heating in large spaces with high ventilation rates.it is important to pay proper attention to the DMRT as a design factor. Vol. D.. January-February 1984. as explained and shown in the equations given. is not a single physical constant for a given room under given thermal conditions. 3. * * Mclntyre. 48-50. 2. 1980. however. both of which showed good agreement. The effectiveness of radiant and warm air heating systems were studied theoretically. In comfort analyses it is important to consider the directional mean radiant temperature as a variable. Pg. of convection. "Warm Air and Radiant Heating: Steady State Power Requirements". D e c . D. A. the author comes to the following conclusions: (1) The simplified calculation method in the CIBS Guide and a more detailed computer program show very good agreement. and of evaporation. 17. such as in factories.1. vertical and horizontal plates is presented graphically as a function of temperature and dimensions.. 1973. the possibility of substantial savings must be set against the increased discomfort due to high levels of radiant asymmetry. * * Prince. J. Higher efficiencies are achieved by conventional electric bar fires (78 percent) and by downward facing panel heaters (up to 80 percent). air pressure should be sufficient to counteract the forces of infiltration. "The Efficiency of Radiant Heat Sources". The author performs an analysis of the effect of air infiltration on temperature stratification. May 1982.-Fred. The radiant efficiency of horizontal cylinders. Measured radiant efficiencies of a selection of commercial heaters are given. and Brailsford. (5) The performance of the warm air system is reduced by temperature stratification if the inlet air temperature is high and the warm air is inadequately mixed. 40. actions can be taken to reduce substantially or eliminate the problem: i) ii) Heating system delivery to the floor should be at least 45 percent of the input.radiant or warm air heating. Vol. Warm-air systems that approach elimination of B-10 . fall below 50 percent. The Building Services Engineer. offer an economical and practical approach to the elimination of stratification. the following. iv) Systems with high infrared output to the floor. "Causes and Prevention of Air Temperature Stratification".. Heaters must be kept as far as possible from the ceiling. A.. Efficiencies of several types of heaters. (4) Radiant heating is more economical than warm air heating in large spaces with high ventilation rates. Feb. R. in both cases. * * Mclntyre. J. * 21. combined with a pressurized air supply (tempered if necessary). Plant Engineering. D. (3) Radiant heating can show advantages when only local warming of part of the room is required. He concludes the following: If present conditions or an analysis of a new heating system design indicates stratification. The increased transfer coefficient found with forced warm air systems will depress their heating performance. * 22. iii) Air must be introduced in the upper portion of the plant or be tempered. described as a radiant. For 8 at 10 range block ft. Limitations of the generalized patterns are discussed and the suggestion offered that the proposed standard panel distributions be used as an initial design arrangement from which to depart in seeking greater simplification. The recommended patterns are: 1. a 12 ft. F. wherever possible spacings outside of the from 9 ft. semi-profile and profile attitudes. spacing as the optimum and 10 ft. ceiling a square pattern with 20 ft. should be avoided. or corrections for such diverse effects as inter-reflections. Generalized patterns for the center lines of ceiling panels have been determined for the standing subject in semi-profile position. ceiling height a square pattern similar to that described above is recommended. * * Raber. but with 14 ft. for panels of usual size in rooms having a ceiling height between 8 ft. 2.. and Hutchinson.s t r a t i f i c a t i o n require substantial expenditure of energy to keep the a i r continually circulated and homogenized. to 15 ft. the use of a single panel in the center of the room is not recommended. or 10 ft. to 21 ft. optimum spacing is recommended (range 17 ft. less than two percent of the energy radiated from a conventional panel passes directly to the occupant. non-diffuse emission and the many other factors which include use of any completely generalized solution for optimum design of all installation. For rooms smaller than 14 ft. Unlike rooms having 8 ft. Use of the basic data to compute the absolute shape factor of a subject in any position with respect to a panel of any shape and size is described and illustrated. B. From those data information is also obtainable as to the fraction of energy received by a prone subject from an elementary ceiling panel. For 10 ft. Basic experimental data has been presented to permit direct determination of fraction of energy received by a seated or standing subject from a wall or ceiling panel of elementary area. for best results. to 14 ft.. 1944. For 12 ft.). ASHVE Trans. W. ceilings. that is. * 23. adjustment for localized exposure. ceiling room does have satisfactory uniformity of direct radiant exchange when the entire ceiling is used as a panel. and 12 ft. the order of magnitude of the shape factor is 1-2 percent. "Optimum Surface Distribution in Panel Heating and Cooling Systems". ceiling height use a square pattern with center lines spaced ft. All experimental determinations are in groups of three based on the position of the subject with respect to the source both as to vertical and horizontal components and for full face. as the limiting range. F. 3. B-ll . x 14 ft. a pattern consisting of centrally located crossing panels running along both axes of the room is satisfactory. Vol. Various types of checks were applied to the calculation in order to verify their validity.24. a computer model was developed for the transient heat transfer within an arbitrary closed room with radiant and convective heat exchange between the panel and room surfaces. 1967.9. W. The conclusions of these authors are as follows: 1) The theory of diffuse gray body radiation exchange. 2. 93. may be taken. constant. ASHRAE Trans. "Configuration Factors for Radiant Energy Interchange with Triangular Areas". These can be used in certain spaces where it is convenient to partition room surfaces into triangular patterns. ASHRAE Trans. V. * * Saunders. "Configuration Factors and Comfort Design in Radiant Beam Heating of Man by High Temperature Infrared Sources". Graphs of configuration factors for triangular areas have been developed and are presented. when modified to include a new parameter called the Beam Utilization Vector. H.0.95.. Vol. In the model. is found to amount to 25% of the direct irradiation of the occupants. the room air temperature was initially set back. J. and. at a m =• 0. 1974. Pt. respectively. Harry J.. J. * 25. Mean body absorptance values for man irradiated by I-R sources having color temperatures of 2500 K (T-3 quartz lamps) and 1200 K(atmospheric gas-fired burners). "The Effect of Wall Reflectivity on the Thermal Performance of Radiant Heating Panels". ASHRAE Trans. heated by radiant panels. provides a rational basis for calculating the incident radiant flux density and energy absorbed from a directional source by a man standing or sitting in any source-man geometry. P. and at 0. 2. and conductive heat transfer through the walls. 1. or comfort level. when clothed. The abstract for this work is as follows: "Numerical results are presented for the heat loss from a room. 73. or ratio of mean directional to mean diffuse radiant intensity in the solid angle connecting a high intensity plane-point radiating source with an absorptive surface. Rapp. 1987. 3. * * Sauer...65 and 0. in the specific case cited. while the mean radiant temperature was adjusted with the radiant panels to keep the operative temperature. A. Pt. The results suggest that walls reflective in the infrared can reduce the steady state heat loss to the ambient by B-12 . respectively. as determined by reported reflectance measurements on white skin and clothing. . For these calculations. when wearing bathing attire. 2. and Andrews. Pt. as a function of wall reflectivity. 80.8 and a m . * 26. for purposes of design. A method for calculating the effect of heater energy re-radiated from the floor is presented.. George and Gagge. however.. Specific analytical and numerical consideration is given to radiant interchange in cylindrical and conical B-13 . no more than 10-15% greater than using the weighting factor approach. A method of analysis has been devised for determining the radiant interchange among surfaces. S. Jnl. having a low emissivity in the infrared and high emissivity in the visible spectrum (0. Some paints and wall papers are available with these properties. It is shown that by careful selection of a solution algorithm for the heat balance equations. Experiments should be undertaken to determine the actual savings in a test room and an occupied home. * 28. F. yet the walls radiate at higher equivalent blackbody temperatures since radiant heat from the panels is reflected by the walls. and Walton. Vol.addresses the problem of calculation of building space sensible cooling or heating load from sensible cooling or heating load from sensible heat gain or loss in an efficient manner on a digital computer. E. of Heat & Mass Trans. G. the former method can be made much more efficient than suggested by early computer codes which used it. both in terms of cost. The formulation uses and generalizes the exchange factor concept (which was initially devised for specularly-reflecting surfaces) and the radiosity concept (which was initially devised for diffusely-reflecting surfaces). since requirements are similar to the requirements for selective surfaces needed for solar collectors (Agnihotri and Gupat 1981. This paper ." They also stated that: "Preliminary calculations showed that IR-reflective walls have the potential for significant reductions of steady-state heat loss.75 microns) to give a normal appearance. Various forms of the analytical method are presented that are suitable either for overall engineering-type computations or for more detailed local investigations. 1. Since there is a loss of flexibility and accuracy with weighting factors. will likely depend upon user acceptance. comfort and aesthetic appeal of the wallpaper or paints used. it is shown that over-all computation times for this method thereby achieved are. E. M. These wall coverings should be selective surfaces. 1980.• * 27." -. Moreover. but appear normal in the visible spectrum. * * Sowell. Int. and Lin... referred to as the heat balance method and the ASHRAE weighting factor method. Vol.. the heat balance may be the preferred approach in the future energy analysis computer programs. 8.35 to 0. Two widely used methods are considered. Future work may be needed to develop inexpensive wall paints or papers that are reflective in the IR. ASHRAE Trans. "Efficient computation of Zone Loads". "Radiation Heat Transfer at a Surface Having Both Specular and Diffuse Reflectance Components". 86. * * Sparrow. Pt. and more might be developed using information gained from research on solar selective surfaces. L. The success of this scheme. 1965. each of which may have both specular and diffuse reflectance components. for many important cases. N.reducing the wall temperatures. ASHRAE Trans. 92. It would also be possible to do calculations of room loads and temperatures as they are affected by the operation of the air distribution system. 3. In this paper. B-14 . 86. George N. In general. The algorithm will make it computationally practical to account for previously neglected effects such as nonlinear. Pt. it is found that the radiant efflux from a cavity increases as the specular component becomes a larger fraction of the surface reflectance. G. Pt.. a single circulation loop forms. M. W. ceiling.. It is proposed that the radiant interchange in a room can be adequately modeled by assuming that each surface radiated to a fictitious surface which has an area.cavities and to radiant transport through a circular tube. and Low. 2. the following conclusions were drawn: 1) The overall flow circulation in a room reverses direction between heater "on" and heater "off" cycles. * * Walton. 2. During extremely cold outdoor conditions or during intermittent heater operation. and temperature giving about the same heat transfer from the surface as in the real multi-surface case.. ASHRAE Trans. the typical velocities. This approximation leads to a more accurate heat balance and can be used for large numbers of surfaces without greatly increasing computation time. "Airflow in Rooms with Baseboard Heat: Flow Visualization Studies". the airflow pattern changes and includes two circulation loops. "A New Algorithm for Radiant Interchange in Room Loads Calculations". 2. * 29.. D. 1986. it is shown that the radiation interchange algorithms of the NBSLD and BLAST loads programs can be significantly improved. V. D. The warmest air passes from the heater directly to the center of the room to enhance mixing and reduce temperature stratification. Temperature stratification can be significant. 1980. A similar statement applies for the transmission of radiant energy through a tube. Results are presented for various subdivisions of the surface reflectance into specular and diffuse components. The basic flow pattern demonstrates relatively high air velocities near the walls. emissivity. In both cases. vol. and floor with stagnant air in the room's center. A model study has been performed to visualize the airflow that occurs in a room with baseboard heat. The full field flow pattern. and temperature distributions in the room are reported for three different test cases. nonconstant interior convection coefficients and heat conduction between simultaneously simulated rooms. From the results of those tests. A.. * * Spolek. * 30. Herriot. and (3) forced convection between the building and its external environment (such as wind-driven ventilation through windows. •* 2. The heat transfer data base from which the correlations were derived was generated by a validated numerical-analysis computer program.. "Convective Heat Transfer in Buildings: Recent Research Results" ASHRAE Trans. "Correlations for Convective Heat Transfer from Room Surfaces". Numerical simulations of wind-driven natural ventilation are presented. -k * Bauman. pt. The convective processes investigated in this research are (1) natural convective heat transfer between room surfaces and the adjacent air. It is shown that such differences can have a significant impact on the accuracy of building energy analysis computer simulations.. ASHRAE Trans. Altmayer. The convection coefficients presently recommended by ASHRAE are internally inconsistent and in disagreement with recent research results. which is a possible method for general calculations. The correlations are applicable to a class of room configurations with cold and warm surfaces on opposite vertical walls. Interzone coupling correlations obtained from experimental work reported in this paper are in reasonable agreement with recently published experimental results and with earlier published work. The ASHRAE coefficients can give significant errors. Convection Coefficients 1. R. More accurate correlations for B-15 . There are substantial differences in the heat flows from the various methods. F. the numerical analysis results have been produced by a computer program based on a finite-difference scheme. 1983. F. In particular. A. and Kammerud.. doors. Gadgil. future research needs are suggested. E. The experimental work has been performed on small-scale water-filled enclosures.. have been developed. Results obtained at Lawrence Berkeley Laboratory (LBL) for surface convection coefficients are compared with existing ASHRAE correlations. F.. 89.. (2) natural convective heat transfer between adjacent rooms through a doorway or other openings. .B. 2A. Altmayer. based on empirical and analytical examinations of convection in two-dimensional enclosures. M. Vol.. Gadgll. . pt. Bauman. The authors compare their correlations for convection coefficients with the ASHRAE coefficients for vertical and horizontal surfaces. 89. The correlations extracted from this data base express the heat transfer rate in terms of boundary conditions relating to room geometry and surface temperatures. and differences of as much as 50% are observed. A. the transition to turbulence for convection in enclosures occurs at a Rayleigh number about one order of magnitude larger than is generally accepted. Nansteel. 1983. E. Vol. Correlation of the rate of heat transfer from room surfaces to the enclosed air. They exhibit good qualitative agreement with published wind-tunnel data. Recent experimental and numerical studies of convective heat transfer in buildings are described. S. A complicated correlation description is given. and important results are presented. C . and other openings). Kammarud. Finally. J. R. 1A. This means that a laminar flow correlation is applicable to a much wider range of Rayleigh numbers than previously recognized. Danter. W. . E. and Witzell. W. The final section considers the overall steady-state heat balance on a room and presents the heat balance equation in a form which provides the framework for developing the analysis of non-steady conditions. * *. The analysis is used to examine the effect of re-interpreting the room temperature in the conventional approach as the arithmetic mean of air and mean surface temperature. Feb. Some equations are given. E.. The index temperature so defined is the 'environmental temperature' of the Guide. The main purpose of this paper is to give an account of the considerations underlying the 'environmental temperature' introduced in the 1970 edition of the IHVE Guide.. ASHRAE Journal. A systematic analysis of the problem shows that the most consistent approximations are obtained in a revised representation of the convective and radiation exchanges in which the heat transfer within the room is related to an index temperature defined as a weighted average of air and room surface temperature. ASHRAE Journal.. This approach shows a less consistent performance and although it can in many cases provide a good approximation it cannot be generally relied upon to do so. Values of the configuration factor or heat transfer coefficient may be determined for parameter values not shown by the construction of a simple cross-plot or interpolation. B-16 . * 3.. "Free Convection and Radiation Heat Transfer from Cylindrical Fins". J . Fontaine. Vol. E. * * C o l l i c o t t . 1965. It originates in a re-examination of the problem of approximating to the effect on the thermal response of a room of the radiation heat exchanges within the room.. 1963. Conventionally these are dealt with interns of a combined radiation-convection transfer between room air and surfaces but the underlying assumption is an over-simplification. Dec. "Heat Exchanges in a Room and the Definition of Room Temperature". W. 0.convection coefficients are needed because they have a significant impact on predictions of building energy consumption. H. Dec. Fontaine. The Building Services Engineer. E. and Grosh. "Radiation and Free Convection Heat Transfer from Wire and Tube Heat Exchangers". H. E. This i n v e s t i g a t i o n i s concerned with a cylindrical fin protruding from a constant temperature source into a constant temperature environment. 41. 1974. Appropriate weights are one-third and two-thirds respectively.. and presented graphically as a function of the diameter to spacing ratios of the wires and the tube and GrPr. * 4. . * * Collicott. R. Effective radiation configuration factors and free convection heat transfer characteristics were determined for wire and tube heat exchangers. * 5. walls --. Eno. * 7. floors Upward 0. * * IHVE.6. ASHRAE Trans. "Thermal and other Properties of Building Structures".1. Pt.. hr ftôF/Btu Building Element Walls Heat Flow High Emissivity Surface (E=0.9) Low Emissivity Surface (E<=0. 1967. P.85 3. W S Kongres . 1985.Building Design and Performance. 2.9 hr ftôF/Btu The following table is presented for inside surface resistances. They are numbers which are very close to the values given in the ASHRAE Handbook of Fundamentals.2. IHVE Guide. The main conclusion drawn from the present work emphasizes the need of a multizone approach of such types of dwellings by computational means. A3. flat or pitched roofs. G.60 1. * 8. Rc. "Combined Convection and Radiation from Rectilinear Fins".0 hrôF/Btu R c . It could be used as a first approximation for finned radiation tubes. Nusgens.. Clima 2000 . It does not consider some of the practical problems of finned tubes. Surface Resistance. The author presents an analysis of the combined effect of convection and radiation from rectilinear fins on tubes. "Convective Heat Transfers within a Large Open Plan Office Area: Experimental Results for Dynamic Buildings Simulations". Vol.. J. Heat exchanges were estimated and parametric studies performed by computer simulations..W S Messe.72 Ceilings or roofs. * •* Hannay. such as fin-tube bond and variability of the convection coefficient on the fin and tube..1.05) Horizontal 0.43 hr ft2oF/Btu Rc.7 1. Air movements and indoor temperature measurements in a large open-plan office area have shown quite definite convective couplings between a warm central core and cooler peripheral zones in both natural and HVAC conditions.16 B-17 . upward flow to ceilings -. Liebecq. Burton E.24 Ceilings and floors Downward 0. Vol. downward flow to floors -. 1. 1977. This section of this guide presents various values for convection coefficients for inside surfaces for heat loss/gain calculations. Some of those given are listed below. 73. 0.05 3. relatively still air conditions (no infiltration). the effect of room size is not significant. 2. 3. D. Convection from ceiling: 2..9. Convection to walls: q e = 0. Natural convection data given by other investigators for small heated plane surfaces were found to be in good agreement with all of the equations listed in the first conclusion except the equation for natural convection from a heated ceiling. the following equations apply in calculating natural convection heat transfer at room surfaces. 5.. 4. L. The approximate combined film conductances (natural convection and radiation) for the heated panels based on the difference between panel temperature and the room air temperature for a normal size room with high emissivity surfaces are as follows: hrc = 2 for heat floor panel at about 85 F. and therefore. Heating. In a ceiling-heated space qe . Ohio. Room size in a ceiling-panel-heated room has a significant effect on the unit convection from the ceiling and the floor. "Natural Convection and Radiation in a Panel-Heated Room". In a floor-heated space B. B-18 .25/De 1. Min. F. Cleveland. A. May 1956. and an empty. 3. 1.3 1 / ! ^ 0 * 0 8 2. the effect of room size on the total heat transfer isriotimportant. Convection to floor: Same as convection from ceiling. Piping & Air Conditioning. Convection to ceiling: Same as convection from floor. The data reported in this paper were obtained with the entire floor area or ceiling area used as a heated panel. the interchange factor between the heated panel and its enclosure may be approximated by the hemispherical emissivity of the panel surface. In a completely enclosed space with high emissivity room surfaces. V. G.29 (At) 1 . J. and Vouris. Convection to walls: Same as for floor-heated space.041 (At)^. unlighted room. However. Convection from floor: qc = O ^ ^ t ) 1 . In a floor-heated room. On the basis of research done at the ASHRAE Laboratory. Schutrum. Parmelee.32 /** 0 . Under these conditions the following conclusions may be drawn: 1. the convection heat transfer from a heated ceiling is small in comparison with radiation exchange. T. a uniform environment wherein all surfaces other than the heated panel were at a uniform temperature.. The convection coefficient for small free-edge plates may be 6 to 10 times as great as that for a heated ceiling. C . 200. based on the assumption that the turbulent boundary layer flow is a fair approximation of conditions to be found in practice. D. G. 53. 7. However. The dependence is very strong in the vertical direction. 10.W S Messe.• » "Experimental Study of Radiation and Convection Heat Exchange in Rooms for Energy Analysis Program Models". unheated surface preceding the heat transfer area.Building Design and Performance. and Huebscher. Since turbulence is an important factor in heat transfer.. Parmelee.. The air stream turbulence in the wind tunnel was measured and found to be 1. 6. 11.h r c — 1.000. Pedersen. Leverenz.000 and 1. and natural convection effects at low velocities. The experiments performed so far have produced several interesting results. after corrections were made for the observed effects of natural convection. 3. J. D. A geometric dependence of the local film coefficient has been identified. G. D. 1985.. Most satisfactory agreement with the friction curves was obtained by defining the length as the length of the heat transfer surface. Examples of the application of the results to problems in the field of heating and air conditioning have been given. V. Further study is indicated in such matters as air stream turbulence.5 percent. 2. Spitler. Vol. J. W S Kongres . 4. ASHRAE Trans. Conclusions: 1. J. while not as B-19 .1 for heated ceiling panel at about 120 F. Clima 2000 . 1947. The effects of surface length and air velocity on forced convection heat transfer between a smooth flat plate and a parallel stream of air have been measured for Reynolds numbers between 19.f "Forced Convection Heat Transfer from Flat Surfaces". The data for the completely turbulent boundary layer flow were in substantial agreement with a formula based on skin friction measurements of flat plates and were in reasonably close agreement with similar data to be found in the literature. Vol. and. C.. R. Bunkofske. it is suggested that future tests include a measurement of the general turbulence by an accepted method. it has been pointed out in the literature that the effect of unheated surface preceeding the heat transfer surface can be significant. The data for the laminar boundary layer were also placed in agreement with the analogous skin friction formula. 5. R. 2. strong, there does seem to be a dependence in the horizontal direction. The range of convective film coefficients produced by the various experiments includes the standard ASHRAE coefficient (after subtracting the radiation component) . While it is unlikely that the standard ASHRAE coefficient is correct for all building heat transfer conditions, its use in combination with a well-mixed model seems to be reasonable for the conditions covered in these experiments. An alternative model fo room heat transfer has been considered. Although the question of how an effectiveness model would be implemented remains unresolved, it was shown that, for a given geometry, the effectiveness parameter was highly correlated to mass flow and temperature difference. * 12. * * Wilkes, G. B. and Peterson, G. M. F., "Radiation and Convection from Surfaces in Various Positions", ASHRAE Trans.. Vol. 44, 1938. Results are given for radiation and convection coefficients for practical configurations found in HVAC. Horizontal and vertical orientation is considered, as is the surface emissivity. This appears to be the source of convection and radiation coefficients, which are published in the ASHRAE Handbook of Fundamentals. B-20 C. General 1. Bahnfleth, Donald R., "Symposium: Field Performance of Infrared Heating Systems", ASHRAE Journal, June 1968. This is a short article describing the activities of the ASHRAE Task Group on Radiant Space Heating and an ASHRAE Symposium that was held in 1968. * 2. * * Bak, Richard, "Hydronic Radiant Floor Heating Staging a Comeback", Air Conditioning. Heating and Refrigeration News, March 1985. Discussion of hydronic radiant floor heating systems. Presents the idea of using plastic piping in place of metal pipes. A general discussion of the installation, application and disadvantages is given. * 3. * * Banhidi, L. , "The Thermotechnical Dimensioning of Radiant Strips and Borders for the Heating of Communal Buildings", Building Science. Vol. 9, 1974. This paper describes a special form of radiant strip heating which offers a suitable solution for the supply of housing and communal buildings with hot water and radiant heating, taking into consideration also the comfort criteria. The thermotechnical calculations are based on Kollmar's method, the design parameters were determined by measurements. * 4. * * Blossom, J. S., "High Temperature Water Heats New School", Heating Piping and Air Conditioning. March, 1959. Discussion of a specific application of a split system in a school. It delivers radiant heat from the pipes that transport high temperature water to fan ventilator units. It looks at the details of the specific design for a school. * 5. * * Blossom, J. S., "Pipe HTW through Classrooms", Heating. Piping and Air Conditioning. July, 1964. This article discusses the use of high temperature water pipes through rooms in order to improve comfort. It is the radiant portion of split type of system. It is the discussion of an application for a specific building. B-21 6. Cambel, A. B. , "Model Study of Radiant Heating", Ph.D. Dissertation, University of Iowa, 1950. This is a thesis done in 1950 and a copy is not available from Dissertation Abstracts. Apparently a copy could be obtained from the library at the University of Iowa at a rather high cost. * 7. * * Carroll, J. R. , "Radiant Systems for Heating", New Methods of Heating Buildings. BRI-760, National Academy of Sciences - National Research Council, Washington, D. C , 1960. This is a discussion of the general characteristics of radiant heating systems. A table of emissivity values for building surfaces is given. General descriptions of radiant exchange; low, medium, and high temperature systems; thermal comfort aspects, and spot heating are presented. * 8. * * , "Concrete Code Restricts Panel Heating", Heating. Piping, & Air Conditioning. October 1951. A discussion concerning the possibility of the American Concrete Institute changing their code for structural concrete." such that it would be possible to have embedded pipes for radiant heating systems. * 9. * * Faust, Frank H., "New Electric Heating Systems", New Methods of Heating Buildings. BRI-760, National Academy of Sciences- National Research Council, Washington, DC, 1960. This article discusses electric heating with the greatest emphasis being on the elctric heat pump. It discusses ceiling panels and other electric panels used for radiant heating. The author presents a short discussion on specific benefits, limitations, applications, costs, and construction requirements. * 10. * * Holden, T. S., "Calculation of Incident Low Temperature Radiation", ASHRAE Journal, April 1961. The author discusses methods for calculating low temperature radiation to building surfaces from the sky, the ground and from neighboring buildings. Emphasis is placed on a method developed for the case of a building surface at any angle to the horizontal with the ground at any slope. B-22 11. Hough ten, F. C , et. al., "Heat Loss through Basement Floors and Walls", ASHVE Trans.. Vol. 48, 1942. Experimental results for basement heat losses are presented. Heat flows, as a function of the time of year, are given as are the ground temperatures at various levels. They observed heat loss reduction as the ground temperature increased, and reduced heat loss from the center of the basement floor. * 12. * * HPAC Engineering Data File, "Design and Control of High Temperature Hot Water Heat Consumers", Heating. Piping and Air Conditioning:. June, 1960. This article presents design guidelines for the use of high temperature hot water systems. Part of the use can be by direct radiation to spaces. General design guidelines are given for heat exchanger selection, boiler selection and operation, and control systems. * 13. * * Johnson, T. E. , "Radiation Cooling of Structures with Infrared Transparent Wind Screens", Proceedings of the Second Workshop on the Use of Solar Energy for the Cooling of Buildings, UCLA, August, 1975. Energy conserving radiation cooling schemes for dwellings in high humidity climates have usually failed due to the deleterious effect of the wind. In this paper, the cooling mechanisms at work in wind conditions are examined. A radiator system using an infra-red transparent wind screen that doubles as the structural envelope is proposed and supporting experimental results are presented. A one family dwelling built with these radiation panels can carry 50 percent of the 24 hour cooling load. Worst case conditions give radiator coefficients of performance twice that of exisiting appliances. * 14. * * Kweller Esher, "Criteria for Mechanical Energy Saving Retrofit Options for Single Family Residences", New and Existing Single Family Residences. ACEEE 1984 Summer Study on Energy Efficiency in Buildings. This paper estimates energy savings, and provides performance and selection criteria, for mechanical equipment options for single-family homes; all from prior studies reported in the literature. Performance and selection criteria are presented as advantages, disadvantages and limitations for each option. Four broad categories of energy-saving mechanical options were investigated: space heating, water heating retrofit options, heat pump water heaters, and recovery of central air conditioner waste heat by desuperheaters. Gas- and oil-fueled forced-air furnaces and hydronic (hot water) space-heating B-23 equipment were treated in this report. * 15. * * Morant, M. A. and Strengnart, M., "Simulation of a Hydronic Heating System; Radiator Modelling", Clima 200 - Heating. Ventilating and Air-Conditioning System. Vol. 6, W S Kongres - W S Meese, 1985. The number of elements taken into account for the modelization of the radiator is chosen depending on the interest of the heating system simulation. If we are interested in calculating the integration of the energy delivered by the radiators over a period longer than one hour, a model with only one capacitance is quite good. One should keep in mind that the temperature evolution obtained with a logarithmic At is better than that observed with an arithmetic At, particularly if the emission is calculated for small water flow rates. Moreover, an optimization of the control system is better performed if the radiator is divided in several elements. 0 * 16. * * MacLeod, G. S., and Eves, C. E. , "Baseboard Radiation Performance in Occupied Dwellings", Heating. Piping. & Air Conditioning. February 1950. In order to determine the performance of baseboard radiation in field installations, tests were conducted in several houses of diversified construction. Winter comfort conditions obtained were evaluated by the observation of the interdependent factors of air temperature distribution, room air velocity, mean radiant temperature, and relative humidity, as well as from the comments of the occupants of the houses studied. An outline of field test procedure and a description of test equipment are given as a basis for future work. Their conclusions were as follows: 1. Size, shape or construction materials of the structures had little effect on the overall performance of the baseboard radiation. 2. Air temperature differentials from floor to ceiling and from room to room were less than in houses heated by more conventional systems." 3. Baseboard radiation systems were free from inherent drafts. 4. Indoor relative humidity was observed to be satisfactory without the use of humidification devices. 5. Highly satisfactory results were obtained from the use of simple control systems. B-24 17. Olivieri, J. B., How to Design Heating-Cooling Comfort Ceilings. Business News Publishing Company, Birmingham, Ml, 1971. A textbook on the design of HVAC systems. It contains a very short section on the description of radiant heating systems. * 18. * * Peach, J., "Radiators and Other Convectors", J.I.H.V.E., Vol. 39, Feb., 1972. This paper is concerned with equipment which emits heat by the combined processes of radiation and convection. If the heat exchanger surface is exposed for all to see it is termed a radiator; if enclosed, a convector. Reference is also made to heat emitters in which air is forced over the heat exchanger surfaces by a fan incorporated into the unit. Heat can be discharged into a room from a unit point, along a line or over an area. In common parlance unitary equipment includes radiators, convectors, fan-coil units, etc. Linear equipment embraces skirting heaters and undersill heaters, whilst area equipment would include underfloor heating, electrically conducting paint, heated wallpaper, etc. These three categories can exist in either of two forms: where the heat is either transferred to the air by free (natural) convection or by forced convection. This paper is limited to unitary types under free and forced convection and linear types under free convection. Operation of the equipment on low temperature hot water systems only has been considered. * 19. • * * Pierce, J. D.f "Application of Fin Tube Radiation to Modern Hot Water Heating Systems", ASHRAE Journal, Feb. 1963. Discussion of radiant baseboard systems and how they perform. Only general information is presented. * 20. * * Rapp, George, "Analysis of Free convection and Radiant Heat Transfer in Valance Heat Exchangers", ASHRAE Trans., Vol. 72, Pt. 1, 1966. The specific fluid flow and heat transfer mechanisms present in valance and baseboard heat exchangers have been identified and quantified by application of modern convection and radiation theory. It was found that conditions strongly favorable to stable laminar free convection exist for temperature differences 50 < At < 300F; and that radiation transfer, apparently neglected heretofore, plays an important role at source temperatures of 200 F and higher. Mathematical general equations giving the net radiant flux within and heat balance on any valance exchanger enclosure have been derived and their application illustrated numerically by prediction of the radiation-convection B-25 No. James D. 73-83. 1949. "Electric Space Heating with Active Boundary Members". * * Weigel. The predicted results agree quite well with full size room rating tests on a similar valance when both are compared at the 200 F source temperature level. 25. Institution of Electrical Engineering Proc.5-20% depending on the setback schedule. as compared with the larger area unheated room surfaces. R. thus combining the functions of thermal insulation and low-density electric heating. ASHRAE.. and these are described. Trans..outputs of a specific design at the source temperature of 200 and 300 F. * 21. "Selected Segment Hydronic Heating System". It has shown that neither the valance (at temperatures up to 300 F) nor the lower temperature ceiling is of much importance in control of body heat loss. * * Rickman. W. H. No. A new type of hydronic forced hot-water heating system is described that provides room-by-room temperature control in a series pipe loop system. 55.. V. "Heating a Basementless House with Radiant Baseboard". The importance of the valance vs the ceiling on occupant irradiation and body heat loss was evaluated. and Harris. Energy Conversion Mgmt. Pg. In common residential processes. These panels are mounted over large areas of the walls and ceiling. W. Vol. A summary of the test results of this investiagtion and the conclusions that may be drawn are as follows: 1. 114. and that valance temperatures much higher than 300 F would be required for any worthwhile effect in this regard. * 23. 1985. Experiments show that the system delivers at least 49% more BTU's per foot of pipe length to rooms selected for heating than to unheated rooms. New mathematical models are required to represent such active members. 1967. * 22. 7. The radiant baseboard is particulary adapted to maintaining B-26 . * * Roots. S. The results of earlier papers permit the building of computer models of electric space-heating processes. the analysis showed that 100 F increase in source temperature (from 200 to 300 F) approximately doubled the output. K. ceiling. the boundary members (walls. Calculations show that temperature setbacks of 3-9^ in selected rooms reduce the yearly heat loss by 5.. 1. floor) have hitherto been passive. Vol. July. It operates by alternately pumping segments of hot boiler water and cool return water in a timed heating cycle. and so now many of the boundary members can become active. and so were representable by available mathematical models. A recent development in North America is the wall panel constructed of glass fibre into which is woven a fine mesh of electric-heating wire. and 30 in. The problem of maintaining adequate indoor humidities for comfort in winter cannot be separated from consideration of good building construction. as measured 3 in. were affected. 6. 4. Dirt patterns of this type can be eliminated by limiting the water temperature to a maximum of 200 F. since long. thus maintaining constant room air temperature even while the outdoor temperature was changing rapidly. low units of this type cover a large percentage of outside exposure. 7. Average air temperatures. Temperature differences between the occupancy zone. However.comfortable floor slab temperatures in a basementless structure. The hot-water heating system using radiant baseboard responded quickly to sudden changes in load. water temperatures in the radiant baseboards. 22 percent relative humidity was obtained in the basementless home. 5. and consequent fluctuations in room air temperatures. Faint dirt patterns were observed on some of the walls above the radiant baseboard after nine months of operation. were approximately 70 to 71 F for all indoor-outdoor temperature differences encountered when the temperature at the 30-rin. above the floor. 3. At 10 F outdoor temperature. B-27 . The longer lengths of the on and off periods of the burner and circulator resulted in greater fluctuations of room air temperatures. as measured three inches above the floor. level was 72 F. 2. Fuel consumption was*not affected by setting of the adjustable differential on the thermostat. were only of the magnitude of a degree and a half to two degrees. 41. "The Practical Approach to Heat Loss Calculation". Vol. London. where local radiation intensities are important.D. The authors have developed a new method for calculating radiant exchange by subdividing the surfaces into small equal squares. Berglund. * 4. "Mathematical Models for Predicting the Thermal Comfort Response of Building Occupants". KSU. The size of the squares is dictated by the accuracy one wants in the results. Building Services Engineer. * 2. Comfort Conditions 1. J. * * Bull. Vol. ii. * 3. H. This technique allows local intensities to be more accurately predicted. ASHRAE Trans. the author tries to show. This is applicable in evaluating the exchange of energy between a person and his surroundings. Pt.. March. 1967. which are given in the ASHRAE Handbook of Fundamentals. P. "Heat Transmission by Radiation". ASHRAE Trans. L. 3-28 . Journal of the Institution of Heating and Ventilating Engineers. "Calculation of Thermal Comfort: Introduction of a Basic Comfort Equation". A step by step procedure is presented. and Erkerlens. He suggests. and Fanger.. Pt. 1963. Vol.. Many references are also provided for the development of the comfort equation.. C . 1978. These models are used when simulating a building's thermal behavior and coupling it to comfort conditions. In this paper. T. 1.(2/3) MRT + (1/3) T a i r of calculating heat transmission in buildings gives identical answers to those arrived at through use of tedious heat balance equations. however. 73. that the proper determination of room heat requirements involves the use of resultant temperature as a comfort index and environmental temperature for heat flow. This paper presents the development of the Fanger Comfort Equation and the Fanger Comfort Equation and the Fanger Comfort Charts. * * Fanger. 1974. These models are the Pierce. * * de Heer.. They all use the heat balance equation together with some physiological parameters to predict the thermal sensation of a person in an environment. Larry. It provides much of the data and figures which exist in the current comfort literature of ASHRAE. 0. Three comfort models are reviewed and equations are presented. through simple numerical examples. 84. Jan. that the environmental temperature method T e . .. It contains the following chapters: Introduction. Thermal Comfort . This is a classic textbook discussing all of the parameters of human comfort and how they are evaluated and interrelated. McGraw-Hill Book Co.Analysis and Applications in Environmental Engineering.5. 0. 1972. The Influence of Certain Special Factors on the Application of the Comfort Equation. Fanger. Conditions for Thermal Comfort. P. B-29 . New York. Practical Assessment of Thermal Environments. Thermal Environmental Analysis and References. It is very useful for making calculations and finding references. Radiation Data for the Human Body. Calculation of Mean Radiant Temperature. and (3) Winslow's Skin wettedness Index of "Thermal Discomfort" (DISC) defined in terms of the fraction of the body surface. A new index PMV* is proposed for any dry or humid environment by simply replacing operative temperature T Q in Fanger's Comfort Equation with SET*. as a standard. 41. March. The discussion centers around the definition of the environmental temperature. Building Services Engineer.6. measureable temperatures produced in the room when the rate of heat supply suggested by the calculation has been used. wet with perspiration. E.. Fobelets. defined as the equivalent dry bulb temperature of an isothermal environment at 50% RH in which a subject. "A Standard Predictive Index of Human Response to the Thermal Environment". The author presented these for convective heating systems and in a later paper for radiant heating systems. B-30 . is illustrated for typical HVAC situations and with a new Comfort-Humidity psychometric chart for indoor environments. (2) Fanger's Predicted Mean Vote (PMV) Index. nomenclature and relative numerical values of the various temperatures involved in load calculation for rooms have been variously described and discussed and there is confusion in some of the IHVE Guide calculations. Temperature and sensory indices of human response to the thermal environment are often expressed in terms of the known in a controlled laboratory environment.(2/3) MRT + (1/3) T a i r Corrections are proposed to the definition of T e based on comfort conditions and calculations. Pt. • 7. 2. The use of PMV* as a sensor of heat stress and strain. The author points out that much of the confusion arises from a failure to distinguish clearly between the temperature used in calculating the heat requirements and the actual. 1986. P. while wearing clothing standardized for activity concerned would have the same heat stress (skin temperature Tg^) and thermo-regulatory strain (skin wettedness. w) as in the actual test environment. T e and its definition. P. The nature. Vol. 92. 1974. G.. The IHVE Guide defines this as T e . ASHRAE Trans. The three rational indices of this type to be considered are (1) ASHRAE's Standard Effective Temperature (SET*) Index. A. Gagge. defined in terms of the heat load that would be required to restore a state of "Comfort" and evaluated by his Comfort Equation. required to regulate body temperature by evaporative cooling. . L. * * * Harrison. The classic difference between PMV and DISC as predictors of warm discomfort occurs at very high and very low humidity but both lead to essentially the" same judgment at average humidities (40-60% RH or 1-2 kPa) . A. and Berglund. "Environmental Temperature and the Calculation of Heat Loss". V.. In other standards the requirement is given as an interval for the PMV-index. R. Heerwagen. D. because it takes into account the total dry heat loss from a person. "Comparison Between Operative and Equivalent Temperature Under Typical Indoor Conditions". T. Only then is it possible to minimize the expenses for heating and air-conditioning buildings without sacrificing an acceptable thermal environment. They have applied this to several situations in buildings. but it is only useful for describing the general thermal comfort at air velocities < 0. ASHRAE Trans. W. .1 m/s. The general thermal comfort may be described more correctly and accurately by using an index like PMV or by using the equivalent temperature than by using the operative temperature alone.. air humidity and the equivalent temperature into account. and clothing conductances on comfort and energy requirements. G. * * Madsen. C. 91. V.. "Developing Office Building Design and Operation Strategies Using VWENSOL and the Comfort Routine". Pt. 1985. A... The equivalent temperature is more correct. B. V... clothing. Olesen. it measures the integrated influence of air temperature. Pt.e. Similar calculations were carried out for the effect of exterior envelope on comfort effects and alternate temperatures. L. i. and Varey. It is shown that the use of the equivalent temperature can save energy during summer conditions but also that it can be necessary to increase the temperature during winter conditions in order to keep the thermal comfort at an acceptable level. and air velocity. even at higher air velocities. B. This index takes the activity level. They consider the effect of cold or warm surfaces radiating to the occupant (strong influence). N. In this paper it is discussed whether the operative temperature is satisfactory or the equivalent temperature should be used as a better expression because it also takes the air velocity into account. 86.. J. * 9. mean radiant temperature. Their conclusions were as follows: The operative temperature is a good one to use when evaluating the heating and cooling loads of a room or building. F. In some of the new standards for thermal environments. KIppenhan. 1980. A sensor for measuring the equivalent temperature directly has been described.8. K. IB. Emery. ASHRAE Trans. a certain operative temperature range is given as a requirement for the thermal environment. as well as window size and number of glazing panes (strong influence).. B-31 . 1. Kristensen. This article discusses the development of a "comfort condition computer program" for use with an energy analysis computer program.. 59. 1974.03 to 2.. D.87 to 2. V. * * Michaels. Rotterdam. the means for thermal sensation. and Lebrun.76. October 1984 (NTIS . B. 70. ranging from 3. For college-age females undergoing 3-hr test periods at rest with air temperature at 75 F and floor temperatures ranging from 75 to 100 F. CIB/RILEM Symposium on Moisture Problems in Buildings. J. Nevins.24 for floor temperatures from 75 to 100 F. moved from slightly cool "3" roward an ideal "4" for comfortable. a higher air moisture causes a warmer B-32 . 2. * 12. by both researchers and official bodies. A. where subjects' assessments on different scales during the same experiment apparently disagree. ranging from 2. M. * 11.63. moved away from an ideal "2" for comfortable toward "3" for hot. A. Results for tests with college-age females standing while performing light work also showed significant effects of floor temperature on foot comfort but not on thermal sensation. R. The author's conclusions were as follows: 1.38 to 3.10. College Age Females". Electricity Council Research Centre. "Evaluation of Thermal Discomfort". Sample means for thermal sensation scores ranged from 3. A. This paper examines the techniques which have been employed to assess the degree of thermal discomfort. The paper considers the problems involved in transferring subjective judgements made in the laboratory to the real world. and discusses to what extend field studies of comfort and behavior can contribute. 3. ASHRAE Trans. Based on foot comfort. At the same time. The mean for foot comfort scores ranged from 1. This study found a positive effect of air humidity on thermal sensation: for same operative temperature. Mclntyre... Proceedings of the Second Int. * * Moisan. floor surface temperatures as high as 85 F do not cause serious discomfort when the air temperature is 75oF. Some are internal. Other inconsistencies exist between different experimenters. jG. "Comfort in Damp Cold Air with Radiant Spot Heating". It is shown that there are many inconsistencies. With increasing floor temperatures means for foot comfort scores. 1964.. where apparently similar comfort recommendations have been arrived at by very different reasoning. K. Capenhurst.PB 85-189975). and the criteria which have been used in setting acceptable limits to environmental variables.07 to 3. and Feyerherm. the data show that there exists a statistically significant effect of floor temperature on foot comfort vote. . "The Effects of Floor Surface-Temperature on Comfort Part III. England. sensation. This effect is much higher than that predicted by the comfort equation in the thermal neutrality region and with RH > 67% . But it appears that there is some discomfort directly related to humidity itself. Perhaps other effects would appear if the time of permanency was increased by some hours in high humidity conditions. On the other hand, no other criteria concerning physiological or hygienic aspects have been considered. * 13. * * Nevins, R. G., "Criteria for Thermal Comfort", Building Research. July-August, 1966. A general review of thermal comfort conditions. some of the early references in this area. * 14. * Presents a history and * Olesen, B. W., Mortensen, E., Thorshauge, J. and Berg-Munch, B., "Thermal Comfort: in a Room Heated by Different Methods", ASHRAE Trans.. V. 86, Pt. 1, 1980. The present experiments have shown that all nine heating methods investigated are able to create an acceptable thermal environment in a well insulated room with one frontage including a double-glazed window exposed to steady-state winter conditions (outside temperature down to -5oC, and air infiltration rates up to 0.8 air-changes/h). When the temperature level in a room provides thermal neutrality (PMV=0) for sedentary person near the frontage, there will be only a small likelihood of local discomfort and the thermal conditions will be acceptable in the entire occupied zone. Only with a radiator at the back wall did the predicted percentage of dissatisfied (PPD-value) at a position near the radiator increase significantly from the optimal value (from 5 to 12%) . In all tests, vertical air temperature differences, radiant temperature asymmetry and floor temperatures were inside established comfort limits. There was a risk of mean air velocities higher than 10 cm/s along the floor in the occupied zone nearest to the frontage when the down-draft along the window and from the air infiltration was not counteracted by an upward convection from the heating system. In general, the highest measured air velocities were in the test with the two floor heating systems (approx. 15 cm/s). * 15. * * Ronge, Hans. E., and Lofstedt, Borje E., "Radiant Drafts from Cold Ceilings", Heating, Piping, & Air Conditioning, Uppsala, Sweden, September, 1957. Contains a statement of methods of study and results of 20 tests made on 5 persons who, after spending from 1/2 to 1 hr. beneath a warm ceiling, moved to beneath a cold ceiling in an experimental room and remained there until B-33 temperature measurements on various body surfaces indicated equilibrium with surroundings. Tests included condition of (1) upper body naked and subject at rest, (2) one layer clothers and subject at rest and (3) 2 to 3 layers of clothes on upper body and subject doing light work. Air temperature covered range from 60.8 to 68 F. Good correlation is shown between skin temperature equilibrium values and mean of air and ceiling temperature. A comfort chart based on the results is included. Experimental results were checked against measurements made in an underground factory room with cool celing. * 16. * * Rohles, F. H., Jr., "Temperature or Temperament: A Psychologist Looks at Thermal Comfort", ASHRAE Trans. . V. 86, Pt. 1, 1980. Five studies are reviewed which address the psychology of thermal comfort. They are summarized as follows: (1) under identical temperature, adding wood panels to the walls, carpet, and comfortable furniture made people feel warmer than when they were in the stark, sterile setting of the room before it was modified; (2) at 65oF (18.3oC),. secretaries who were informed that a radiant heater was operating in the modesty panel of their desks, felt warmer than those who were not informed that it was operating; (3) when people were told that the temperature of a room was 74<>F (23.3<>C) when it actually was 72oF (22.2oC), 70oF (21.1oC), or 68oF (20oC), they were just as comfortable as when the room temperature was 74oF (23.3oC); (4) in a study to determine if comfort was related to the season of the year, it was found that cool temperatures are preferred over warm temperatures in the summer and the opposite is true in the winter; (5) based on a questionnaire in which temperatures were ranked as cooler-than-comfortable, comfortable, and warmer-than-comfortable, a Preferred Comfort Envelope was proposed that ranges from 70oF (21.1oC) to 76oF (24.4<>C). * 17. * * Springer, W. E. , Nevins, R. G., Feyerherm, A. M. and Michaels, K. B., "The Effect of Floor Surface Temperature on Comfort: Part III, The Elderly", ASHRAE Trans.. V. 72, Pt. 1, 1966. Conclusions reached in this investigation were the following. 1. For elderly females and males undergoing 3-hr test periods at rest with the air temperature at 80 F and floor temperatures ranging from 75 to 100 F, the data show that there exists a statistically significant effect of floor temperature on foot comfort vote. A statistically significant effect of floor temperature on thermal sensation did not exist for elderly males but did for elderly females. With increasing floor temperatures, the means for foot comfort ranged from 2.12 to 2.57 for the elderly males and from 2.25 to 2.64 for the elderly females. At the same, the thermal sensation votes ranged from 4.28 for the elderly males and from 4.04 to 4.60 for the elderly females. 2. With an air temperature of 80 F, the data show that floor surface B-34 temperatures as high as 85 F did not cause serious discomfort for elderly females or males. 3. With a floor surface temperature of 90 F the individuals are apparently experiencing a discomfort due to the inability to reject heat readily via conduction and radiation, but at a floor temperature of 95 F the greater stress has induced the body to dissipate a greater portion of the heat via moisture evaporation thus providing lower thermal sensation and foot comfort votes. Increasing the floor temperature to 100 F overcomes the advantage of increased evaporative losses and once more higher thermal sensation and foot comfort votes result. 18. Stevens, J. C. , Marks, L. E. and Gagge, A. P., "The Quantitative Assessment of Thermal Comfort", Environmental Research. Vol. 2, 1969, pp. 149-165. Discomfort aroused by lowering or raising the operative temperature of a subject's environment was found to follow the "power law" that governs many dimensions in the domain of"sensory psychophysics. To a first approximation, discomfort caused by cooling grows as the 1.7 power of shifts downward in temperature from the level that feels comfortable; discomfort caused by heating grows as the 0.7 power shifts upward from the level that feels comfortable. One group of 8 subjects matched numbers to the degree of discomfort (magnitude estimation); another group of 12 subjects adjusted the loudness of a white noise to match the discomfort (cross-modality matching). These verbal and nonverbal methods gave approximately the same result with regard to the quantification of thermal discomfort. B-35 E. Thermal Comfort-Radiant 1. Albert!, M. and Rugger, R. , "A Method to Check Thermal Comfort Conditions in High Industrial Buildings Provided with Radiant Panels Heating Plants", Clima 2000 - Indoor Climate. Vol. 4, W S Kongres - W S Messe, 1985. A procedure was developed using the Fanger Comfort equations for determining the PMV and PPD indices in radiant heated, high ceiling industrial buildings. The method uses computerized procedures to evaluate the thermal environment at any position in an industrial building. This technique and calculation is applied to the Italian heating energy savings regulations. Very few details are given on how the equations are solved, what convection coefficients were used or what emissivities were used. Sample results of PMV are given as are some of the resultant surface temperatures. * 2. * * Bahnfleth, D. R. , "Physiological Effects of High Intensity Radiant Beam Heating", ASHRAE Jnl.. Nov., 1964. This is a description of ASHRAE activity in radiant heating during the early 1960's. It also discusses ASHRAE Research Project-41-Physiological Effects of Radiant Beam Heating. * 3. * * Baker, Merl, "Improved Comfort Through Radiant Heating and Cooling", ASHRAE Journal, February 1960. A description of the advantages of panel heating and cooling are presented along with how this affects the sensation of comfort. The article is not very in-depth. * 4. * * Barihidi, L. , Dr., Somogyi, A., Kintses, G. , Besnyo, J., "About Local Discomfort Effects Caused by Asymmetric Radiation Occurring During Winter in Dwelling Houses", Clima 2000 - Indoor Climate. Vol. 4, W S Kongres - W S Messe, 1985. This article deals with the laboratory analysis of a special case of a symmetric radiation that commonly occurs in everyday life. Namely, in winter what are the acceptable surface temperature boundaries for the inner side of a main front wall, in relation to the local discomfort feeling of a person sitting next to the wall? Sitting face to face with a cooled panel brought about in both cases of gradual and random temperature change more unfavorable reactions (20% discontent) of thermosensitivity than when the participant sat with back to the cooling wall (temperature limits were 22oC, 20oC and 16oC for facing with continuous temperature, facing with shorter random temperature B-36 periods, and back to wall respectively). * 5. * * Berglund, L. G. and Fobelets, A. P. R., "Subjective Human Response to Low Level Air Currents and Asymmetric Radiation", ASHRAE Trans. V. 93, Pt. 1, 1987. "The responses of 50 subjects wearing winter clothing (0.86 clo) to twohour-long exposures of various kinds of winter indoor conditions were studied. The conditions included air speeds between 0.05 and 0.5 m/s (10 and 100 fpm) and asymmetric radiation to a cold wall that produced radiant temperature asymmetries ranging from 0. to 20 K (0 to 36 F). The study was done at neutral or preferred temperatures and at conditions 3°C (5.A F) lower. Some of the conclusions are: The operative temperature concept for combining air and mean radiant temperatures into a single temperature scale is an effective means of characterizing and controlling complex environments, although the coefficient A in the operative temperature equation of ASHRAE Standard 55-81 may be too low at high air speeds. -• The neutral operative temperature, calculated according to ASHRAE Standard 55-81 from the experimentally determined neutral conditions, for velocities of 0.25 m/s (50 fpm) or less were unaffected by radiant temperature asymmetries of 10 K (18 F) or less. Thermal acceptability at neutral conditions was unaffected by air speeds of 0.25 m/s (50 fpm) or less and RTAs of 10 K (18 F) or less. Thermal acceptability decreased when radiant temperature asymmetry increased beyond 10 K (18 F) . Thermal acceptability decreased when air speed increased from above 0.25 m/s (50 fpm) even at neutral conditions. An operative temperature 3°C less than neutral is probably too low for human sedentary occupancy as thermal acceptance of such conditions was only 63% in this study. There were differences in the subjective responses between the men and woemen of this study. The perception of draft was a linear function of air speed and temperature and independent of radiant temperature asymmetry. The sensation of local cooling was related to RTA and independent of air speed. There was no interaction between velocity and RTA on the subjective responses of this study. That is, effects from velocity and radiant asymmetry are independent and additive. Relationships were found relating thermal sensation with thermal preference, comfort, and thermal acceptability." * 6. * * Berglund, L. G., Gagge, A. P., and Banhidi, L. J., "Performance of Radiant Ceiling and other Heating Systems Controlled for Equal Comfort with an Operative Temperature Sensor", Proceedings of the Third International Conference on Indoor Air Quality and Climate. Stockholm, August, 1984. B-37 A typical office space was heated separately by (1) radiant ceiling panels, (2) forced air, (3) baseboard and (4) floor heating systems. Each system was controlled with the same "operative temperature" sensor, whose set point was held constant at 23<>c. Twenty subjects experienced each environment for 3 hours. The comfort and whole body thermal sensations were not statistically different for all four systems, although air and radiant temperatures and other thermal characteristics differed widely. The present experiments demonstrate that operative temperature, as a single control input, is an effective way to regulate heat for comfort in complex environments. * 7. Berglund, L. G. Conditions Baseboard, V. 91, Pt. * * and Gagge, A. P., "Human Response to Thermal Maintained in an Office by Radiant Ceiling, Forced Air and Floor Heating Systems", ASHRAE Trans., 2, 1985. A typical office space was heated separately by (1) radiant ceiling panels, (2) forced-air, (3) baseboard, and (4) carpet heating systems. Each system was controlled for equal comfort with the same "operative temperature" thermostat whose set point was unchanged throughout the test series. The office was contained within an environmental chamber. The steady-state power consumption per unit floor area for all systems averaged 9.1 Btu/hft^ (98 W/m^). The radiant system used the least power. Air temperatures in the occupied space were most uniform with the floor heating system. Twenty subjects experienced each environment for three hours. Periodically, subjects indicated their thermal sensations for whole body, head, and feet, local discomfort, comfort, and whether the environment was thermally acceptable or not. Comfort and whole body thermal sensations were not statistically different for all four systems and thermal acceptability averaged 94%. Though subjects indicated their feet were slightly warm with the carpet heating, they preferred this system to the others tested. The experiments demonstrate the usefulness of operative temperature as a control parameter for complex environments, such as those produced by radiant ceilings. The thermal environmental characteristics of the four heating systems made it difficult to test each system at exactly the same comfort level. However, the wall-mounted operative temperature sensor-controller, whose set point was unchanged, maintained the comfort, thermal acceptability, and thermal sensation responses of the occupants in the test space at similar, if not identical, levels. The air temperature in the occupied zone was most uniform with the floor heating and the least so with the baseboard system. It was also very uniform with the forced-air system. The difference between air and globe temperatures was the largest, with the radiant ceiling system using the least power. The floor system used the most power. In these tests, the forced-air system caused the most local discomfort, followed closely by the radiant ceiling system. The occupants reported the least local discomfort with the floor system, even though the mean thermal sensation for their feet was slightly warm. The occupants' preference ranking of the four systems from preferred on down was: floor,, baseboard, forced-air, and radiant. From the thermal sensation, comfort, acceptability, and local discomfort responses, one would have expected the radiant to be preferred over the forced-air system. The radiant panel system used 89% of power per unit floor area as the B-38 Boyar. From the experimental results. The conclusions of the authors follow. 4. June 1966. G. "Effect of Floor Heating on Man's Comfort and Thermal Sensations". There was also a surprisingly low floor temperature with the forced air system. floor surface temperature. of Delaware. system used 116%. 1985. The baseboard system used 109%. The purpose of this study was to show the thermal effect of the posture in a room during sitting on the floor or sitting in a chair. and Gagge. it was found that the case of sitting on the heated floor was more comfortable than sitting in a chair. Some temperatures are presented from some experimental results for a forced air system and a radiant ceiling panel system.3 clo) would be necessary. humidity control is rather unimportant for comfort as is air movement below 30 fpm. Operative temperature is approximately the average of the air and mean radiant temperatures present. 18-20oC. and the carpet heating Berglund. Heating. Comparing the results of this experiment with other experiments indicated that the thermal environment was slightly warm. the air and mean radiant temperatures are seldom equal. "Thermal Comfort and Radiant Heat". The comfort conditions for floor heating found in this study are as follow: air temperature.Indoor Climate.. E. W S Kongres W S Messe. In passive solar buildings.forced air-system. Univ.. P. An environment expected to be thermally acceptable to 80% or more of its sedentary occupants would have operative temperatures between 68 and 80oF. Proceedings of the Third National Passive Solar Conferencer American Section of the International Solar Energy Society.. "The Influence of Radiant Energy Transfer on Human Comfort". L. Techniques for predicting the mean radiant temperature from the expected surface temperatures of the space are clearly described by Fanger. R. A.1 m/s. Piping and Air Conditioning. under 0. 26-28oC. K. The radiant ceiling panels provided higher MRT near the glass surfaces than did the forced convection system. 9. air velocity. 10. Bohgaki. B-39 . Clima 2000 . The author shows the variations in MRT that can be expected from different methods of providing heat (forced air or radiant) to a space. Of course over this large temperature range appropriate clothing adjustments (from 1 to . Vol. Fortunately for the passive building. Comfort conditions in such buildings can be conveniently described in terms of the operative temperature of the environment. 1979. It was concluded that the floor temperatures recommended here were undertaken in accordance with those accepted in current practice. in accord with current practice. A. The paper describes a series of experiments on five subjects. Measurements of the skin temperatures of the subjects' feet were made and the results showed that subjective reactions of the men and women to thermal stimuli were very similar. did not modify the conclusions' as to the maximum desirable ceiling temperatures reached in the earlier tests. (2) Studies of localized sensations of warmth have enabled design criteria B-40 . Ancillary observations made in the field are included in a discussion of the relation between the experimental results and existing British practice. 1964. The paper describes experiments designed to obtain quantitative assessments of the risk of producing discomfort in rooms with heated floors. Proceedings of Symposium of the Institution of Electrical Engineers in London. March. and it is concluded that the recommended temperatures are. Chrenko. "Radiant Heat and Thermal Comfort". Discomfort was closely associated with the floor-surface temperature and with the temperature of the skin of the sole of the foot. which give maximum desirable surface temperatures of panels of various dimensions embedded in ceilings of different heights. The subjects wore their usual clothing and footwear.. A. and such differences as were found between the responses of the two groups of subjects at various floor temperatures were due to differences in footwear. F. based on the results of these experiments. on the whole. A. * * Chrenko. London. The authors conclusions were stated as follows. the causes of which are discussed. designed to obtain quantitative assessments of the risks of producing unpleasant conditions in rooms with heated ceilings and includes tables. April. April. The experiments were carried out in the laboratory on five men and three women who were (a) sitting and (b) walking about. A method of computing the mean radiant temperature at a point is given. The results of a separate series of experiments investigating the effects of a cold wall on persons exposed to radiation from a heated ceiling. * 12. "Heated Floors and Comfort". F. was most closely related to the elevation of mean radiant temperature at head-level due to the heated ceiling.11. Journal of the Institution of Heating and Ventllatine Engineers. 1955. • * * Chrenko.. 1952. Journal of the Institution of Heating and Ventilating Engineers.. F. The risk of discomfort. London. * 13. (1) Existing psychophysical scales of warmth are of no use in designing new systems of radiant heating or the assessment of old systems. Electricity and Space Heating. "Heated Ceilings and Comfort". and further experiments on a group of 150 subjects. 1970. The purpose of this study was to determine experimentally geometrical radiation data necessary for calculation of the radiant heat exchange between humans and their environment. 0. showing the percentage of people feeling discomfort due to overhead radiation. infrared heating systems. . * * Fanger. B-41 . This data is necessary when analyzing panel heating or cooling systems. A curve has been established.. 0. * * Fanger. P. Angelius. Olesen. Pt. 0. 76. .. It is recommended that a heated ceiling should not provide a radiant temperature asymmetry exceeding 4<>C in spaces with high standards for the indoor climate. The corresponding limit for the ceiling temperature can be found from a figure for different sizes and heights of the heated ceiling. L. 86. . The curve applies for sedentary people who feel thermally neutral for the body as a whole. Clothed subjects were exposed to radiant asymmetry from walls and floors for 3 1/2 hours. "Radiation Data for the Human Body".. P. 0. * 15. "Radiation and Discomfort" ASHRAE Journal.. Increasing discomfort due to increasing overhead radiation with lowered air temperature.to be given in ordinary physical terms. as a function of the radiant temperature asymmetry. This study presents results that indicate that people are not particularly sensitive to asymmetric radiation from surfaces. and Langkilde. In order to have 10% dissatisfied occupants it requires from 7 to 25oC radiant temperature asymmetry. G. B.. and Kjerulf . can be attributed to warmer head and colder feet. "Comfort Limits for Heated Ceilings" ASHRAE Trans. pt. Vol. 2.Jensen. It shows that in practice the limits will rarely be exceeded. (3) Studies of the sensitivity of the skin to radiant heat have yielded results which have practical application in design. P. and effects due to cold surfaces. Barihidi. * 14. Vol. A useful diagram relating percent dissatisfied to the angle factor yields the allowable temperature difference between the air and wall. * 16. ASHRAE Trans. Many figures are given for specific geometries. P.. February 1986. * * Fanger. 2. Less than 5% of the population are then predicted to feel uncomfortable due to overhead radiation. 1980. W. . A constant level of comfortable operative temperature implies either that there is thermal equilibrium with the environment or that there is a constant but small rate of body heating or body cooling. Briefly stated the conclusions of RP-41 are as follows: 1. A. 71. Pt. G. which can B-42 .. 4. P. A normal physiological state is not necessarily the most comfortable to an individual. * * Gagge. A..7 to 20 microns.. . "Mean Radiant and Operative Temperature for High-Temperature Sources of Radiant Heat". M. October 1964. ASHRAE Journal. evaporative loss and the vascular regulation of peripheral blood flow. 1965. M. and Rapp. * * Gagge. D. D. A method has been presented showing how much radiant heat may be required to balance out the discomfort of low ambient air temperatures. Vol. comfort may be described by the operative temperature. such as metabolic rate. but usually falls in a wide range of operative or ambient temperatures. For environmental conditions with varying ambient temperatures and radiant heat. A practical level of operative temperature for comfort useful for this method is 80oF (unclothed) and 72<>F (clothed). The objective of this paper was to outline a method of standardizing measurements of high intensity radiation by any radiometer in order to derive a mean radiant temperature (MRT) and an operative temperature (T Q ). Physiological factors.. * 18. J. can each affect the level of thermal equilibrium and perhaps comfort by 2 to 4 degrees fahrenheit.. 3. 1968. The general aim of the project has been the development of a comfort standard for high temperature sources of thermal radiation in the spectral range 0. P. ASHRAE Jnl. * 19. A method has been presented. A. Comfort is not described by any single temperature level. "Exploratory Study on Comfort for High Temperature Sources of Radiant Heat"..17. April. J. ASHRAE Trans. comfort may be described by the operative temperature. G. Hardy. 2. 5. showing how much radiant heat may be required to balance out the discomfort of low ambient air temperatures. and Hardy. . Rapp. Gagge. "Final Progress Report: RP-41-Physiological Effects of High Intensity Radiant Beam Heating". P. For environmental conditions with varying ambient temperatures and radiant heat. A practical level of operative temperature for comfort useful for this method is 80 F (unclothed) and 72 F (clothed). 2. The results indicate greater B-43 . and humidity were held constant. Forty subjects in two experiments experienced conditions in which ceiling temperatures varied between 26. . J. M. air velocity. For high temperature radiant heat. The ratio of the ERF to the environmental constant represents the environmental temperature change caused by the radiant field. effective radiant field (ERF) has been introduced that has useful application to radiant heating (and cooling). . Raising the air temperature did not increase sensitivity to overhead radiation. ASHRAE Journal. but ceiling temperature significantly affected warmth assessment. 16 (4). comfort and physiological response are directly proportional to ERF and to the difference between mean radiant temperature and ambient air temperature. I. In the study. The subjects appraised the environmental conditions by use of a 34-item semantic differential questionnaire.describe with greater accuracy man's thermal response to his environment. and Hardy. the radiant energy received by the unclothed subject from a high-temperature source has been evaluated without specifying its quality by the increase in his sweat rate. "Subjective Response to Overhead Thermal Radiation". and environmental temperature. The effective radiant field is defined as the radiant heat exchanged by an occupant with his surrounding environment when his black body skin (or clothing) temperature is hypothetically equal to the ambient air temperature. The experimental conditions have been so chosen that the increase in the subject's radiation load was balanced by an equal increase in his evaporation loss. and Mclntyre. the re-radiation from warmed floors or ceiling. D. mean readiant temperature. G. Human Factors. The MRT and T 0 . Baldness and seat height were unimportant factors. Air temperature. used here to describe any radiant environment. The sum of this change and the ambient air temperature describes the operative temperature by which comfort and discomfort may be judged in accordance with known physiological and comfort standards. and raising the ceiling temperature did not cause discomfort. A new hypothetical variable. subject baldness.. "The Effective Radiant Field and Operative Temperature Necessary for Comfort with Radiant Heating". D. Conditions of higher ceiling temperature were perceived as cooler than those with the same mean radiant temperature and lower ceiling temperatures. 1967. Rapp. ERF is a summative term and may include the radiant heating from a high temperature source. * 20. Thus. P. * * Gagge.. Experimental variables consisted of two levels of seat height. have been defined in terms of the radiation (hr) and convection (hc) coefficients that would have applied under corresponding conditions of equal wall and air temperature. the ERF is equal to the actual heat absorbed by the occupant from the radiating source. * * Griffiths. * 21. A. it has been possible to describe a complex radiant environment in terms of a temperature scale well associated with everyday experience. For constant ambient air temperature. and the cold radiation from window surfaces. 1974.5oC and 45©C. Discussions concerning the physiological properties of skin.snesitivity to radiant exchange with walls than with the ceiling. "Radiant Heating and Comfort". 1972.. * 24. D.. there were no significant differences between conditions. It seems that the standard on poor evidence where satisfactory evidence.4 yr) experienced three environmental conditions of equal predicted subjective warmth. 6. "Physiological Effects of High Infrared Heating". It also seems that there is no evidence of a preference for a radiant rather than a convective environment and that those designing installations should decide between radiant and convective systems on their individual merits. 1973. standard deviation 6. and Mclntyre. * * Griffiths. I. mean radiant temperature 23.3©C. This supports a previous finding with young men as subjects.5 yr. in environmental temperature formula should be T e = 0. Environment Research.56Ta + 0. that radiant and warm-air environments are not perceived differentially and also suggests that the relative importance of air and mean radiant temperature for warmth is not affected by age. I.44Tr if the air velocity is equal to or less than 0alm/s. Fifty-six women over 55 years of age (mean age 67.. November 1962. D. June. B-44 . A.0oC. * * Griffiths. D. Caution should be used when infra-red radiant energy is used for concentrated heating. "The Balance of Radiant and Air Temperature for Warmth in Older Women". 40.8 C). but different mean radiant and air temperatures (air temperature 26.2oC. as related to radiant energy. V. * 22. A. exists and that the something like sought to do is to evaluate the evidence which temperature and comfort. The Building Services Engineer. air temperature 23. After 40 minutes. and that European upper limits for ceiling temperature are unduly restrictive. mean radiant temperature 26. * 23.7»C. exposure subjects rated the environment on a number of subjective scales.9 C. D. * * Hardy. What the authors have exists concerning radiant recommendations are based fact. air temperature 19. and Mclntyre. James D. mean radiant temperature 17. are given. ASHRAE Journal. and Lorenzi. The authors conducted tests for five days each on the effect of floor heating conditions and ceiling heating conditions.. At a room center black body temperature of 75 F. One particular building wall construction. ASHRAE Trans. R. * * Herrington. Under these conditions the gradient between extremities and environment is increased about 28 percent over comparable heating with ceiling panels operating at 95 F. which conforms to the ASHRAE Standard on Energy Conservation in New Building Design (90-75R). 2. P. October 1949. 1981. "Heating the Perimeter Zone of an Office Building: An Analytical Study using the Proposed ASHRAE Comfort Standard (55 -74R)". The purpose was to find what effect the location of the radiant heating panel had on the human body. Then. as indicated by head temperatures (an average of cheek. and dorsal neck area) and mean exposed skin temperatures and clothing surface temperatures. the authors conducted experiments recording these variables under comparable conditions. Three different types of heating systems were analyzed: baseboard convection. pt. 87. J. Vol.25. 3. 2. radiant floor panels operating at 79 F produce a detectable increase in the temperature of the clothing surface of the lower extremities. all-air. was evaluated. upper hair surface. Physiological considerations are reported which support the view that floor temperatures above 75 F are not desirable... B-45 . an analytical study was performed to demonstrate the dependence of operative temperature on outdoor temperature. Heating. An alternative solution would be the use of a wall section with less glazing area. * 26. other surfaces and air temperatures being within 3 deg of this value. and radiant panel. Gordon H. Measurements were made of the effect of various temperatures of floor surface. Their conclusions were as follow: 1. Foot temperatures under comfortable conditions requiring no house heating are about 10 deg above shoe surface temperatures which are near 74 F when floors are at 71 F. L. & Air Conditioning. The results of this study should be useful in designing and operating office buildings with uniform comfort in the perimeter zones. "Effect of Panel Location on Skin and Clothing surface Temperature".. room center and ceiling surface on subjects. One of the primary conditions of thermal comfort is a skin temperature ranging from approximately 80 F on the toes and sole of the foot to approximately 95 F on the trunk and certain facial areas. with an overall average for the skin surface of 90-92 F. Since room comfort is closely related to the surface temperature of an occupant's skin and clothing. Using ASHRAE Standard 55-74R along with fundamentals of radiant and convection heat transfer. Piping. The results suggest that it would be necessary to raise the air temperature in the space to a slightly higher value to maintain a constant operative temperature. mean comfort votes and mean temperature votes were taken. Hart. * 27. Clima 2000 Indoor Climate. G. Pt. J. * * Lebrun. 1941. This is a discussion about a set of experiments relating radiation to comfort conditions. The duration of exposure was 3 1/2 hours and subjects had different qualities of clothing. The Predicted Percentage of Dissatisfied is computed from every distribution of internal temperatures.. Some information is given about comfort experiments with subjects. For a "badly" insulated building (with large. performed to confirm the physical diagnostic of the inside micro-climates realized by all the heating systems previously studied. single-glazing areas). "Comfort Limits During Infrared Radiant Heating of Industrial Spaces".. 4. * * Houghten. .. What is presented is a rational interpretation of differences observed between experimental and theoretical values of the overall heat transfer coefficient of the exposed wall as of the volumetric heat loss coefficient. ASHVE Trans. "Radiation as a Factor in the Sensation of Warmth". and Sue in. Gunst. air velocity and mean radiant temperature.. Thirty-two subjects were exposed to overhead radiant heating from gas fired infrared heaters. Vol. a satisfactory comfort is only achieved by radiators located below the windows or by warm air. No effect of wearing a helmet was found. The efficiency of the system relates to thermal comfort conditions as well as to energy consumption. this location for a radiant panel i s preferred to the floor location. B. "Thermal Comfort and Energy Consumption in Winter Conditions -. Jan. F. Relationships between percentage of dissatisfied and radiant temperature asymmetry were established for people exposed to infrared heaters. 1979. and some were standing and some seated.W S Messe. Radiant heating systems by floor and/or ceiling panels are examined by the detailed measuring of the inside microclimate in an experimental room in relation with heat transmission through an exposed wall and with ventilation enthalpy flow. . It considers such factors as wet and dry bulb temperature. Jean J. Accepting 5% feeling uncomfortable. and Marret. vol. C. The authors develop a relationship between MRT and effective temperature. The results are compared with those previously obtained with a warm air system.Continuation of the Experimental Study" ASHRAE Trans. •*• 29. W S Kongres . Gunnarsen. 1985.. S. These limits are less restricitve than existing recommendations and standards. L. * 28. 85. Changes in MRT dictate changes in effective temperature resulting in no change in feeling of warmth. Mortensen. heating. Dominique J.4. 2. * * Langkilde. a radiant temperature asymmetry of 10 to 14HC was found permissible.. Since c e i l i n g location of the radiant panel does not produce a significant effect on head temperature. N. . but this second solution is much more energy consuming B-46 . the maximum installed load Pmax(W) is related to the heated area A(m squared) by Pmax . The v. Electricity Council Reasearch Centre.r. Cold radiation on the back and warm radiation on the face are apparently disliked.r. above the recommended limit of 10K. "Overhead Radiation and Comfort". * 31. "Overhead Radiation and Comfort". From this evidence. termed the v. it is shown that asymmetric thermal radiation characterized by a vector radiant temperature of greater than 10K produces noticeably non-uniform conditions. A.. Five experiments are summarized. radiant heating i s as comfortable as the other systems. A vector radiant temperature of 20K does not increase the mean discomfort vote. because of the i n t e r n a l isothermy i t could produce. A measure of asymmetry has been developed. In practice this sets an upper limit to the heat emission that may be provided without increasing the v. This simple relation should be used at the design stage to check the B-47 . warm a i r heating would probably become the most energy saving one. This result is applied to the design of ceiling heating systems. Batiste Publ. * * Mclntyre. and makes recommendations on the maximum loading of ceiling heating systems which can be used without risking complaints of discomfort. January 1977. this level of radiation is noticeable. and hence to the heat output. D. * 30. For a "very well" i n s u l a t e d building. This series of experiments has investigated the reactions of people to overhead heating. For a "well" i n s u l a t e d b u i l d i n g . and may be blamed for causing discomfort.t. Ltd. England. Pg. * * Mclntyre. May 1976. is related to both the size and temperature of the heated area. The Building Services Engineer. D.) is a measure of asymmetry.r. Vol. However. This report summarizes the work on overhead radiation.t. and may be thought of as the average surface temperature of one half of the room minus that of the other. A.r.t. It is recommended that for normal indoor situations a v. The Environmental Section at ECRC has conducted a series of experiments aimed at an understanding of the effects of thermal radiation on comfort. of 10K be regarded as the upper limit.t.t. but buildings with poor insulation and large windows may require further insulation to reduce the power loading. For rooms of normal height.(due to the increase of i n t e r n a l h e a t exchange coefficient) . Capenhurst. 44..700 + 95A. n e v e r t h e l e s s there i s no miracle: radiant heating does not allow a s i g n i f i c a n t lowering of a i r temperature and corresponding v e n t i l a t i o n loss. (NTIS-PB 277 428). and the published results of other workers.r. 226-34. There is evidence to suggest that the direction of the radiation is important. generally from a heating panel extending over the ceiling. There is no difficulty in meeting this criterion in well insulated buildings. and is likely to lead to complaints by occupants. V e r t i c a l a i r temperature p r o f i l e s are a l s o given.. The vector radiant temperature (v. so that perceived warmth was constant across the conditions. This is unlikely to produce any adverse effect on comfort at conventional levels of illuminance below 4000 lux. D. A. One hundred and forty-eight subjects each experienced one of four levels of overhead radiation. * * Mclntyre. * 32. A. Scales of general evaluation showed a slight improvement with increasing asymmetry. However. * 33. and reduced to compensate for the raised ceiling temperatures. The degree of asymmetry is characterized by the vector radiant temperature (v. * * Mclntyre. The concept of the field is probably unfamiliar to most workers in the field of comfort or thermal physiology. However. 1976. 1977.. Any installation using 1000 lux or more is liable to produce uncomfortable conditions. It appears that people are ready to attribute discomfort to unusual aspects of the environment. 287-296. London. exposure. up to a maximum ceiling temperature of 45oC. Tungsten filament sources produce considerably more radiation. "The Thermal Radiation Field". where it is more usual to deal in terms of energy exchanges. "Sensitivity and Discomfort Associated with Overhead Thermal Radiation". Fortunately.. and it is to be hoped that some of the measures. This paper has presented a comprehensive view of the thermal radiation field. this level did not actually increase discomfort. Fluorescent lighting produces an irradiance of about 8 W/mz per 1000 lux. pp. 3.r.). 5.. the vector radiant temperature. * * Mclntyre. 3. 1974. V. several differential radiometers are already on the market. D. D. No. After 15 min. will be used when specifying the thermal environment. Vol. Air and wall temperatures were held equal to each other. = 10K is therefore suggested as a design criterion. the four levels were 0. A. 9 and 14 K. Illuminating Engineering Society. at 70 W/m^ per 1000 lux.r.t. 8. No. 20. Ergonomics. Building Science. particularly the radiation vector and its equivalent. A maximum asymmetry of v.t. 9. * 34. the field concept provides a simpler and more powerful way of describing the thermal radiation environment. and the measurement of the radiation vector presents no B-48 . V. The thermal radiation received from a lighting system may be estimated from a knowledge of the lamp type and the illuminance. a scale which asked whether the hot ceiling caused discomfort showed a steady increase in discomfort with increasing asymmetry. the subjects rated the environment on seven scales. but was noticeable and in practice levels greater than this are likely to produce complaints.acceptability of a system. "Radiant Heat from Lights and Its Effect on Thermal Comfort". 3oC per 1000 lux. Active (relative air velocity = 45 fpm): a. "The Relative Effects of Convection and Radiation Heat Transfer on Thermal Comfort (Thermal Neutrality) for Sedentary and Active Human Subjects". 74. This may be compensated by reducing the air temperature 0. = 1. the level at which discomfort occurs has not been established. * 36. A linear relation was obtained between total thermal radiation and illuminance.3 0 fpm): a. 41. The thermal radiation from a number of different light sources was measured as a function of the illuminance. Pt.51 Females (metabolic rate = 301 Btuh) . D. The constant of proportionality varied from 0*07 W/m2 per lux for tungsten filament lamps to 0. = 1. determined by the ratio of the convection heat transfer coefficient (hc) to the radiation heat transfer coefficient (h^).4 2. C . * 35. The following shows the relative influence of convection and radiation heat transfer. J.435 in.37 c...43. "Thermal Radiation from Lighting Installations". E. 1968. The Building Services Engineer. P. Males and females combined (metabolic rate = 741). 2. The typical modern fluorescent lighting installation of 1000 lux will produce no comfort problems from thermal radiation. A. V. b. V. which implies that the thermal irradiance may be predicted from illuminance.7oC per 1000 lux. ASHRAE Trans. * * McNall. It is still not possible to obtain a reliable instrument for measuring the mean radiant temperature and this fact more than any other has delayed the understanding and acceptance of the importance of the radiation field in determining comfort and warmth.2. Males (metabolic rate = 389 Btuh).008 w/m2 per lux for fluorescent lamps. Radiant" asymmetry per se will not cause discomfort at illuminances up to 15000 lux for fluorescent or 1700 lux for filament installations.59 clo in equilibrium with environments that have a partial pressure of water vapor of 0.4 B-49 . which implies that the increase in mean radiant temperature due to radiation from a fluorescent lighting installation is up to 0. The figure is 0. => 1. recommended value 1. and Schlegel.problem.006 w/m2 per lux for low pressure sodium lamps. Hg: 1. 1973. April. recommended value 1. Sedentary (relative air velocity = 2 5 . he/ti£ = 1. for people wearing clothes with an insulation value of 0. Males and Females combined (metabolic rate = 345 Btuh). * * Mclntyre. V. The regression planes. The results of the statistical analyses performed on the votes of thermal comfort of sedentary male and female subjects wearing clothing with an insulation value of 0. it is felt it applies for less severe exposure to heated panels.20 to a wall 20 F cooler than the balance of enclosure surfaces and thermal sensations of subjects exposed to uniform enclosure surface temperatures belong to the same regression plane.59 clo in equilibrium with environments with a partial pressure of water vapor of 0. P. Therefore the "Thermally Neutral Zone" developed in an earlier study for enclosure surfaces of uniform temperature is applicable for environments of the former type.435 inches Hg and air velocity of 20-30 fpm indicate that: 1. B-50 . 4. Although the previously mentioned "Thermally Neutral Zone is not applicable for thermal environments of the former type. 76.20 to wall panels at 130 F experienced significant discomfort which was found to be caused by the asymmetry of the mean radiant temperature. 2. Jr^ and Biddison. E.5 higher than expected in the case of the 130 F wall. ASHRAE Trans. including the validity of Fanger's equation for predicting thermally neutral environments.. 37. 3. McNall. R. Pt.20 to a wall at 130 F and for subjects exposed to uniform enclosure surface temperatures were found to be significantly different.12 to ceiling panels at 50 and 130 F and radiation shape factors of 0. .The results show generally good agreement with Fanger's comfort equation In the environments investigated. 1970. "Thermal and Comfort Sensations of Sedentary persons Exposed to Asymmetric Radiant Fields". relating thermal sensation with air temperature and mean radiant temperature.. producing thermal sensation votes about 0. The thermal sensations of subjects exposed with radiation shape factors of 0. The same ratio for sedentary and active subjects is felt useful for engineering purposes.shape factors of 0. 1. Thermally "neutral" subjects exposed with radiation. developed for subjects exposed with radiation shape factors of 0.20 to wall panels at 50 F experienced no significan discomfort which could be attributed to the symmetry of the radiant field. Thermally "neutral" subjects exposed with radiation shape factors of 0. E. * * Nevins. Heating Piping and Air Conditioning. seated at rest. effective temperature. 1957. floor temperature and relative humidity. M. comfort data were obtained by subjecting college age students. Correlation coefficients were calculated for the correlation of comfort vote with air temperature. seated at rest and exposed to the test conditions for 3 hr. indicate that thermal sensations of both male and female college-age subjects are not seriously affected by floor surface temperatures as low as 60 F. Foot temperatures recorded at the end of the 60-min exposure indicate that 88 to 91 F is the maximum foot temperature for comfort under the conditions of these texts. The radiation heat transfer in a uniform environment (air temperature equal to MRT) is approximately 25 to 30% of the total.. operative temperature. In addition. Oct. the temperature and the location of the man... A total of 108 male and 21 female students were used. For a 60 F floor temperature. 0. Pt. G. Foot comfort votes indicated that both male and female subjects objected to a floor temperature of 60 F.. R. During the 1957 tests. Floor surface temperature over a range of 65 to 95 F were found to have a negligible effect on the comfort vote when the air temperature was 75 F. 2. G.38. * 39. and Flinner. air temperature 75 F. a floor temperature of 65 F may be too cool for female subjects. The influence of the floor temperature varies with the size of the floor. V. B-51 . roughly correlated with foot comfort vote. floor surface temperatures of 100 F may reduce the radiant heat transfer to 20% of that occurring in a uniform environment. Limited tests involving cold floors. measured on the bottom and top of the foot inside the shoe. Subjects were clothed. 149-153. p. "Effect of Heated-Floor Temperatures on Comfort". For the first phase of a study to determine the effect of floor temperatures on comfort. male and female. and Feyerherm. 73. using male subjects. "Effect of Floor Surface Temperature on Comfort Part IV: Cold Floors". The comfort vote did not show a significant correlation with floor temperature or relative humidity. Foot skin temperatures. air temperature plus mean radiant temperature. Based on theoretical calculations. Nevins. the heat transfer may increase to 150 %. The coefficients show that the comfort vote correlates with those parameters in which air temperature is a predominant factor. R. It was concluded that 95 F is the maximum floor temperature for comfort under the conditions of these tests. ASHRAE Trans. it was found that floor temperature of 100 F significantly affected the comfort vote whereas floor temperatures of 80 to 95 F did not. to various floor-panel-heated environments for periods of 60 min. 1966. A. A. . Sample means the 4. 1968.09 to 2. V. Pt. V. 3. A simplified method of calculation for evaluating the thermal indoor environment of the design stage has been presented. ASHRAE Trans.95.36. * * Schlegel. 70. for example). The mean for college-age males standing while performing significant effects of floor temperature on for thermal sensation scores stayed close to temperatures up to 95 F. It uses the operative temperature. The mean for 100 F foot comfort scores ranged from 2. C. A. In the asymmetric MRT tests. ranging from 3. floor surface temperature and radiant temperature asymmetry.. Based on foot comfort. 1. At the design stage.43. Results for tests with light work also showed comfort. K. * * Olesen. and outdoor climate. P. Pt. Michaels. A statistically significant effect of floor temperatures on thermal sensation and foot comfort does exist for college-age males undergoing 3-hr test periods at rest with air temperature at 75 F and floor temperatures ranging from 75 to 100 F. At the same time the means for thermal sensation. moved away from an ideal "2" for comfortable toward "3" for hot.. B. means for foot comfort scores. ranging from 2. the MRT was separated up to 12 F from air temperature.40. one entire wall of the test chamber was cooled or heated approximately 12 F different from the B-52 . "The Effect of Asymmetric Radiation on the Thermal and Comfort Sensations of Sedentary Subjects". 89. R. * 42. and Feyerherm. and humidity. surface temperatures. The air velocity was held constant at 25-30 fpm and the vapor pressure of moisture in the air was maintained at 0.55.0 mark for floor was 4. "A Simplified Calculation Method for Checking the Indoor Thermal Climate". For the symmetric MRT tests reported previously.435 in. Several large computer programs can be used to predict heating and cooling loads. and McNall. M. 1964. 1983. 2.08 to 2. 74. Hg (45% relative humidify at 78 F. it is desirable to be able to predict the indoor thermal climate that will result from a given combination of building construction.64 to 3. "The Effect of Floor Surface Temperature on Comfort-Part 1. indoor air temperatures. B. heating system. E. Experiments were conducted by exposing sedentary subjects for 3 hrs to environments of symmetric and asymmetric MRT that were in and around the thermally neutral zone. * 41. moved from slightly cool "3" toward an ideal "4" for comfortable. ASHRAE Trans.. J. W. G. II. College Age Males". . 2B. V. Nevins.. but these programs are often difficult and expensive and are used mainly for large buildings. floor surface temperatures as high as 85 F do not cause serious discomfort when the air temperature is 75 F.. With increasing floor temperatures. ASHRAE Trans. The experiments employed only low temperature radiation near room temperature levels. The conclusions are: 1. A. ceiling. even for the relatively mild temperature differences employed in these tests. R. and that at ceiling surface temperatures of up to 60oC angle factors of up to 0. it cannot be concluded that sensations of comfort or neutrality are sufficient to rule out harmful effects that may exist for unilateral cooling. T. where the air temperature was limited to 21oC.12. V. B-53 . It was found that at ceiling surface temperatures of up to 50oC patients suffered no additional discomfort with angle factors. pp. They do not apply to infrared sources. This article considers the effects of radiant heat on the heat dissipation from the human body. of up to 0. since the subjects used in the present tests were not allowed to participate in repeated exposures. and floor. Spangler. and the subjects were seated relative to that wall such that a sphere in the position of each subject would have a shape factor of 0. However. Use of these methods has made climate control practical in applications that would have otherwise been impossible.31 to the heated surface. An experimental survey of the limitations placed by patient comfort considerations on the size and surface temperature of infra-red ceiling heating panels in a hospital ward is described. and of up to 0. Air Conditioning. 43. were permissible. BuildinE and Environment. and Rae. 1977. Various methods are also presented for reducing radiation effects and mean radiation temperature. 143-146. "Industrial Climate Control Versus Radiant Heat". where the air temperature was limited to 23©C. as evidenced by the radiometer readings of mean radiant temperatures which accurately predicted the subjects' thermal sensations. Heating and Ventilating.. Smith. based on a parallel planes measure. 3. 44. Jan. 12. M. 2.balance of the walls. "Patient Comfort and Radiant Ceiling Heating in a Hospital Ward". A chart and basic equations are presented for calculating these effects along with the necessary adjustments in ambient air temperature required to maintain the same relative degree of human comfort..31. The surface temperatures were adjusted so that MRT was equal to air temperature for a sphere in the position of each subject. . No significant discomfort was noticed by the subjects due to the asymmetric MRT of the magnitudes tested. 1965.2 with respect to that wall. The seated subjects could be approximated as spheres for mean radiant temperature calculations. A. A two-sphere radiometer developed by Honeywell was used to measure the mean radiant temperatures. Industr. At very low sensation levels. Stevens. Med. ASHRAE Trans. the error in the use of one globe at 45 in. "Subjective Warmth in Relation to the Density. above the floor will generally give a very B-54 . With increasing sensation level.t. reference point should be placed in the region of the knees rather than the trunk when a man is seated with his legs near a heated panel.62 C and 1. Attention has been drawn to the fact that in non-uniform environments (1) the mean radiant temperature varies throughout the occupied space. Differences between m. and (4) the heat load on a man is dependent on his orientation towards the various radiating surfaces.45.17 C for the standing and sitting models respectively. While the mean of the estimates of m. 1. the proximal stimulus for subjective warmth cannot be skin temperature per se. and Marks. Hence it seems that in indoor environments warmed by any of the more usual forms of heating installation but with no intense radiation from sharply localized sources. 1965. 3. 1970. are reasonably consistent with the data on warmth. one involving a two-layer receptor system. Pt.r. area becomes a progressively less effective determinant of warmth than flux density. E. The exact size of the exponent of the power function depends on the duration of the exposure and its areal extent. Duration.r. Vol. Joseph C.. 1. and as a result.s obtained with the thermopile at 45 in. (3) errors will arise in assuming that the mean radiant temperature at a point is equivalent to the mean radiant temperature as its affects a man. The conclusions of the aughor are quoted below. Two theories. 58. at higher sensation levels it takes a much larger percentage change in area to offset a given percentage change in density. the other a single receptor system possessing properties of adaptation. B.t. Flux density and areal extent can be traded for each other to preserve the same sensation of apparent warmth. 22.. 2. Subjective warmth aroused by infra-red irradiation of the skin grows in magnitude as a power function of the flux density of the irradiation. above the floor was only small. "Assessment of Mean Radiant Temperature in Indoor Environments". * * Tredre. the simple estimation of mean radiant temperatures with a globe thermometer at 45 in. The conclusions from this experimental investigation were as follows. Subjective warmth correlates well with flux density but poorly with flux duration. above the floor were slightly greater but still did not exceed 0. 76. Lawrence.. the rule of trading is virtually complete reciprocity. A number of other common hypotheses concerning the nature of the proximal stimulus are also at odds with the data on apparent warmth.s from two or three globe thermometers gave good agreement with the results obtained from sitting or standing metal models of men respectively. The results suggest that the 45 in. and Areal Extent of Infrared Irradiation". Brit. * 46. Since skin temperature depends on both density and duration. J. V. (2) the average surface temperature is a dubious index of mean radiant temperature. satisfactory indication of the heat load on a standing or sitting man. I 1 B-55 . "Methods for Testing Hydronic Floor-Heating Systems". Electricity and Space Heating. b) Downward losses for different tube spacings and depths. H. London.. 1964. Most of the systems have been tested with the B-56 . Discussion of floor warming in order to take care of loads occurring at a later time. Proceedings of Symposium of the Institution of Electrical Engineers in London. and numerous measurements have been made. S. "Electric Floor Warmings in Commercial Buildings". 91. 14 w/ft^ with a charge time of 12 hours in 24 will produce floor surfaces up to 85 F. Oct. F. 2.. Proceedings of Symposium of the Institution of Electrical Engineers in London. 1953.Research.. March.Some Design Data".5 w/ft^ for an intermediate floor is too high for comfort conditions. J.. W. "Off-Peak Floor Heating . For the German standard DIN 4725. "Floor Panel Heating . V. ASHRAE Trans. 4.. c) Emission of panel . hollow-pot floor.. Design and Development: Some Controversial Factors". methods have been developed for testing the thermal performance of hydronic floor-heating systems. pre-stressed plank floor. Data are presented relating to: a) Surface temperature for different tube spacings and depth of the tube in the concrete floor.from 1. E. hollow concrete beam floor. Electricity and Space Heating. Grammling. Suggests that the recommended value of 22.F. 1964. 1985. Billington. The use of the network analyser to study some probleems of floor heating is described. Pt. It will produce floor surface temperatures in excess of 85 F. The author discusses several systems that can be used for warming floors. H. Journal of the Institution of Heating and Ventilating Engineers. March. N. This was extremely sensitive to tube spacing and depth. These are. Faithfull. and cast in situ floor. He discussed the advantages of warmed floors for comfort and heating.2 to 0. He suggests that in some cases. 2. Bruce. FLOOR PANELS 1. 3.5 Btu/ft* hr HF. different infiltration rates. Pt. * 5. B-57 . Ph. No. Jr. The steady state design water temperature appears to be more than adequate for transient operation. The panel configuration that was considered consists of hot water pipes buried in either a bare or a covered concrete slab floor with a concrete footing and a perimeter insulation. the panel thermal resistance. * •* Hogan. Heat Transfer Analysis of Radiant Heating Panels . 92. Roy Edward. and Blackwell. 1.with an area-weighted average unheated surface temperature. The ASHRAE panel model is acceptable for the geometry considered even though it does not represent the panel heat loss mechanisms correctly.Hot Water Pipes in Concrete Slab Floor. The results of the t e s t s show that measured values of performance differ significantly from figures published in the literature or company catalogs. . Jr. 1986. Particular parameters of interest are the downward and edgewise heat loss. A transient simulation of the panel performance over a typical winter day is presented and a control system is discussed. * 6. 2. The ASHRAE design recommendations are conservative because both the downward and edgewise heat loss and the panel thermal resistance are overestimated. The numerical model is described in detail the results are compared to prior experimental data. The numerical results agree in trend with the prior experimental results. E. The ASHRAE design recommendations for radiant floor heating panels are reviewed and evaluated using the results of a numerical model. Both bare and covered panels are considered. "Comparison of Numerical Model with ASHRAE Design Procedure for Warm Water Concrete Floor Heating Panels". ASHRAE Trans. The ASHRAE design recommendations for radiant heating panels are reviewed and evaluated using the results computed from a numerical model. the panel thermal resistance.. and the required mean water temperature. Isotherms are plotted for the temperature field in both the panel and the earth. R. Particular areas of interest are the downward and edgewise heat loss.. Vol.so-called place apparatus. The ASHRAE design recommendations are adequate and slightly conservative for designing both bare and covered radiant floor heating panels with no infiltration and an AUST equal to the room air temperature.D.. These design recommendations are conservative because both the downward and edgewise heat loss and the panel thermal resistance are over estimated. and an AUST not equal to the room air temperature. The numerical model is described in detail and uses a finite control volume based solution method. The ASHRAE design recommendations are shown to be adequate and are slightly conservative for designing both bare and covered floor heating panels with no infiltration and. Further studies could be made for other panel geometries. August 1979. I t i s therefore clear that exact performance measurements under controlled thermal conditions are necessary for the planning and optimal operation of floor-heating systems. AUST. and the required mean water temperature. B. . equal to the room air temperature. * * Hogan.. Results of the numerical model were compared with prior experimental results and agree qualitatively. . Huebscher. M. 1950. Their conclusions were: 1. Piping & Air Conditioning. The experimental studies confirm the fundamental theory given in another paper. . and (7) position of the pipe grid between the slab surfaces. Simple equations are given for establishing the division of heat flow between the two surfaces. Four in. developed to verify and extend the range of thermal test results. V. B-58 . (2) thermal conductivity of the slab. (5) slab thickness.. * 8. L. E. Hulbert. are compared with values predicted from theory.. * * Humphreys. V. (6) tube outer diameter. A set of curves applicable to the case of uniformly spaced pipes buried within a solid slab having isothermal surfaces has been developed for the purpose of relating the following quantities: (1) rate of heat release (or pick-up) per linear foot of pipe. It is particularly important that the panel be separated from the foundation by suitable insulation. (3) difference between the pipe and mean slab surface temperatures. V. H. An electrical analogue. F. Heating. and Locklin. H. April 1950. is described.. C. Nottage. April. and some details of instrumentation are given..7. * * Humphreys. * 9. Heating. Schutrum. M. as determined by thermal tests and electrical analogue methods. C. Heating. (4) spacing between adjacent pipes. B. W. The greatest part of the heat loss from a floor panel occurs around the perimeter of the panel. C. thermocouples and heat flow meters were installed under four houses to study the heat losses from floor panels to the earth. Heat flow rates and temperature distribution. . F. Nottage. During the summer of 1915. G. C. Equipment for studying heat flow within concrete panels is described and results of tests on three panels are reported. It is in this area that insulation will prove most effective. Schutrum.. "Field Studies of Heat Losses from Concrete Floor Panels".. Franks. Data derived from analysis have received experimental confirmation. R. "Heat Flow Analysis in Panel Heating or Cooling Sections". Its use as an edge insulation was not studied. 3... January 1951. hollow clay tile does not appear to be any more effective as an insulation under a floor panel than an equivalent thickness of gravel. L. B. Pining & Air Conditioning.. Franks. Different kinds and amounts of insulation were placed under the floor slabs. and Franks. The results of the tests in these four houses during the 1949-50 heating season are reported. 2. D. "Laboratory Studies on Heat Flow Within a Concrete Panel". L. C. Piping and Air Conditioning. as in kilns. Heating. . D. "Losses from a Floor-Type Panel Heating System". Sebal. Cunningham. 369. 2. Mills. Edge and rear losses were of the order of 30 percent of the total energy supplied so it may be concluded that insulation is necessary with floor-type systems if economical operation is desired. This paper presents the results of an investigation on heat losses from two floor-type panel heating systems during the 1948-49 and 1949-50 heating seasons. A formula is derived for the estimation of the heat loss through a floor standing solid on the ground and surrounded by a wall. 2. 22-128. Institute of Fuel Journal. Numerical data is given and many photographs of installation procedures are given. Based on the steady-state two-dimensional finite element analysis carried out in the present study.. H.. Hutchinson. L.. * * . furnaces and driers. L. Pt. one depending on the wall thickness and the other on the shape of the floor. 141-1565. E.No. * * Perry. experimental results are also presented giving the actual rating of a unit area of floor panel and the combined film coefficient of heat transfer for such a panel as evaluated from tests on the actual system. H. "An Analysis of Heat Losses through Residential Floor Slabs". The heat loss coefficient is virtually independent of outdoor air temperature and varies only slightly with the deep soil temperature. S. . and Scesa.10. ASHRAE Trans. Le Procede Calendal: de chauffage par rayonnement a' basse temperature". For an unisulated slab about 60% of the total heat loss occurs through the region lying within three feet of the slab edge.. p. W. March 1975. the following conclusions can be drawn: 1. Reuve de 1'aluminium. -. . It involves two constants. 438. and La Tart. "Heat Loss Through a Solid Floor". In addition to the loss data. Piping & Air Conditioning. the latter results are of particular interest in that they permit conclusive decision as to the applicability of the laboratory tests of other investigators to actual field installations. 1985. * 12. * 11. * * Macey. F. T. H. pg. A French article discussing the specific details for installing floor panel heating systems. V. G. B-59 . December 1950. 91. * 13. J.. Vol. The latter dependency has seemingly been ignored in past studies. Gives some design procedure details as well as installation details. R. "Performance of Covered Hot Water Floor Panels . E. Pg. An instruction manual produced by a Canadian (German) manufacturer for hydronic floor radiant heating systems. Maximum temperature should be positively limited to 130oF and preferably 120oF by means of aquas tat cut-off valves on the feed line from the 'blender' to the slab. and Harris. Soil thermal conductivity should definitely be taken into account in any estimate of the slab heat losses since the losses are directly proportional to the soil conductivity and since the latter parameter varies considerably from one soil type to another. Heating. * 16. E. The thermal resistance of both the asphalt time and the rubber tile B-60 . * * Plattis. Pressures should not exceed 15 psi for sustained used at these temperatures. Thermal resistance of the combinations of carpeting and pad ranged from 0. Janca Enterprises Ltd. April. L." * 15. The following is a summary of the results obtained for the test conditions investigated: 1. 1955. 55. "Where Polyethylene Pipe Challenges Metal for Slab Radiant Heating". March 1985. Canadian Builder.87 (F deg) per Btuh (sq ft) for the heavy carpet and 40 oz jute pad. a relationship familiar to the HVAC community. 1963..Thermal Characteristics". * * Sartain.Part I . October.40 (F deg) per Btuh (sq ft) for the rubber pad alone to 1. W. The author states. the constant of proportionality was found to be strongly dependent on both the insulation configuration and the soil thermal conductivity.3." He adds: "Pipe must meet recognized standards.05 (F deg) per Btuh (sq ft).. 3. S. * * . * 14. Apparent thermal resistance of the bare concrete panel was about 1. . Radiant Floor Heating.. Thus. 2. Piping & air Conditioning. "Both laboratory and field experience show that low density polyethylene pipe should be fully suitable for concrete slab radiant heating systems if proper control is exercised. Plasco Manufacturing Ltd. The study confirmed that the losses occur primarily near the edge of the slab and are proportional to the product of slab perimeter and the indoor/outdoor temperature difference. However. accurate predictions of slab heat losses must include considerations of the soil underlying the slab. 5. November. 4. W. L. Glass surface' temperatures measured with floor panel heating were the same as those obtained in the Research Home with conventional radiation.. reverse loss from the panel.was about 0. At design conditions of 80 F indoor-outdoor temperature difference the maximum difference in room-air temperature between the levels 3 in. unless the piping is arranged to permit zoning with the use of more than one water temperature. * 17. covering the floor panels with any type of carpeting had pronounced effects on the water temperatures. Piping & Air Conditioning. B-61 .Part II . The greater the thermal resistance of the floor covering. above the floor. Floor coverings. Covering a floor panel with carpeting did not appreciably increase the seasonal fuel consumption. 5. The AST in Rooms A and B was about 1 F below the room-air temperature while in Rooms C and D the AST was essentially the same as the room-air temperature. 6. S. 3.05 (£ deg) per fituh (sq. Because of the large increase in water temperature required when a carpet is applied to floor panels it may be impossible to balance floor panel systems in which carpeting is used in some rooms only. below the ceiling was 3.Room Conditions". 2. which have a thermal resistance of 0. Cooled air dropped to the floor at the north wall and window and moved in the direction of the south wall with a somewhat high velocity.. At a location 2 ft from the north wall.2 (f deg) per Btuh (sq ft) or less had a neglibible effect on the performance of floor panel systems. 4. and the required boiler size (see Table 4 ) . Heating. and Harris. * * Sartain. Major effects of carpeting over a bare floor panel on the design and performance of a floor panel system are shown in Table 4. The exposed wall surface temperature was about 8 F lower and the AUST was about 4 F lower than the room-air temperature measured at the center of the room 30 in. 7. Addition of floor coverings to bare floor panels reduced the ability of the system to maintain a constant room-air temperature. E. 1956. such as asphalt tile or rubber tile. resulting in poor response. "Performance of Covered Hot Water Floor Panels . above the floor and 3 in. At design conditions.5 F. 8. ft). 1. the greater the resulting room-air temperature variation. the room-air temperature was 3 F lower than at the center of the room. Carpets and pads retarded the flow of heat from the water to the room-air. Heating. 2. The temperatures of the air 3 in. 3. with figures presented in Chapter 12 of THE GUIDE 1953 for heat losses from B-62 . above the floor were practically the same. For a given panel minus room-air temperature difference the panel heat output to Rooms C and D. The measured heat flow from panel to room ranged from 87 percent of the calculated above floor heat loss in Room A to 101 percent of the calculated above floor loss in Room C. . W. 4. January 1954. Comparisons of the reverse loss from the heated slabs. At design conditions. 5. were 15 to 20 percent greater than outputs in Rooms A and B. E. 6. The following is a summary of the results obtained for the test conditions investigated. indicating that there was probably an increase in the rate of transfer of water vapor through the concrete floor slab in room. The effects of the carpeting were to cause an increase in the floor surface temperature along a line toward the center of the room and to smooth out the heat flow profile from the panel to the carpet. S. At design conditions of 80 F indoor-outdoor temperature difference the measured panel output was from 7 to 18 percent greater than the calculated panel output. The reverse loss was roughly twice as great as the estimated heat loss through unheated floor slabs using the heat transmission values given in THE GUIDE 1953. The savings in material and the ease of installation made the vertical insulation the more desirable of the two types. * * * Sartain. Piping & Air Conditioning. which had more severe exposures. "Heat Flow Characteristics of Hot Water Floor Panels". The relative humidity in Room D which had a carpet and pad was consistently greater than that in Room B. 1. It was found that the fuel savings resulting from the use of insulation under the entire floor slab as compared to the use of edge insulation only was too small to warrant the additional cost. Vertical insulation along the inside edge of the foundation wall was as effective as the L type edge insulation. below the ceiling and 3 in.. L. level of 0. with a variation between the floor and 60-in. The air temperatures at the center of the rooms were very uniform.5 deg. and Harris.The MRT as obtained in the center of Room A with a thermo-integrator was the same as the AST for all outdoor conditions encountered. the reverse loss from panels with edge insulation amounted to 20 to 23 percent of the total panel output. . p. 3. The effect on the room air temperature the amount of infiltration air and its the same in the floor panel tests (see shown to be in the ceiling panel tests 4.8 F with a floor panel. Humphreys. the values for total panel output obtained in these tests indicate that the values given in THE GUIDE 1952 (Fig. 2. with no infiltration.unheated floor slabs. 548) are of the right order of magnitude for unheated mean radiant temperature (UMRT) values of around 65 F. . 9. V.5 F deg. level of varying entering temperature was about Fig. and the 90-in. and 71. For a given floor surface temperature and room-air temperature. The authors qualified their conclusions to the fact that these were laboratory tests. level would be 74. one could expect a somewhat higher output per square foot of panel area in an unisulated room or one with large glass area than in a fully insulated room with limited glass area. C. Thus. the room air temperature at the 60 in. * * Schutrum. studies of which are currently under way. Their observations were as follows: 1. G. M.. July 1953.2 F with a ceiling panel. B-63 at the 60-in. . 7. * 19. Parmelee. Room air temperatures were found to be higher for a floor panel heated room than for a ceiling panel heated room for the same room surface temperatures. In none of the tests reported on here did the air temperature gradient in the room between the 2-in. Pending the establishment of data on the radiative and convective components of the total heat output from floor panels. For example. Reference 1). "Heat Exchanges in a Floor Panel Heated Room". 10 this paper) as it was (see Fig. the heat emission rate was much higher near the exposed wall and window than at the center of the room. with an 85 F panel temperature and a 70 F AUST. F. would exceed the fuel consumption when using radiators or convectors by about 10 percent. level exceed 3. the floor panel system had the desirable characteristic of automatically increasing the heat output rate in areas adjacent to points of high heat loss from the room. Heating. 11. Piping and Air Conditioning. L. indicate that at an indoor-outdoor temperature difference of 80 F. the fuel consumption when using floor panels. It was found that while the mean floor surface temperature was uniform across the panel. Lienhard. 1. 1985. H. A description of a panel water distribution system for use in floors. Report No. Air Conditioning. and Shapiro. * * . 1985. One aspect they did not check was the effect of the spatial oscillation in the surface temperature. 2. 3. This would replace the commonly used plastic or metal tubes. V.. * 22. . On the basis of the simulations studies. Slab systems share the advantage of other radiant heating methods in that the air temperature can be lower than for convective-type systems. 91. T. "Underfloor Radiant System Uses 86*i Supply Water". "Analysis of Slab-Heated Buildings". (unpublished) The summary and conclusions from this UNPUBLISHED article are the following. J. H. E. ASHRAE Trans. due to edge and bottom heat losses. 77004. N. 1985. An important design trade-off exists in selecting the depth of the mats. Department of Mechanical Engineering. some general conclusions can be stated regarding the feasibility of slab heating in comparison with other direct heating systems. Placing the mats close to the surface allows for quick response to changing loads. 21. Provides a description of floor panel heating systems and their advantages and disadvantages.20. Shamsundar. Houston. deeper mats cannot respond quickly but are needed to provide enough energy storage in order to use off-peak electric power B-64 .. Analyses of the performance of slab-heated buildings have been presented. but results in inadequate designs. 2. They have shown that the ASHRAE procedure is not only erroneous. Oct. and how the modified procedure can be used to design the amount of insulation needed. "Performance of Polybutylene Pipe in Concrete Heating Panels". They have shown how this procedure can be corrected to increase its accuracy considerably. Alternatively. N. and Tezduyar. H. but it has the disadvantage of allowing the inside air temperature to drop rapidly when the mats are turned off. They have developed computer programs to obtain accurate results and recommend that they be used in critical cases to check the adequacy of the design. * * VanGerpen. University of Houston. The earlier research performed by ASHAVE showed that this needs to be considered only when the tube pitch was larger than twice the slab thickness.. 2. Slab heating actually requires more input energy than conventional heating systems to maintain the same air temperature. Texas. Pt. * 21. J. Heating and Refrigeration News. However. in a given case. 5. 4.and overheating. The main advantage of slab heating is in its ability to use lower cost off-peak power.exclusively. The control of a slab-heating system is difficult because of the phase shift between when the energy is put into the mats and when it is recovered at the slab surface. a complete life-cycle cost analysis would be required to assess the feasibility of slab heating. This can lead to chronic under . B-65 . * * Baker. C. Fortunately. the visible portion of the radiation spectrum is highly reflected while the infrared. 2. This paper presents a simplified theoretical analysis of the removal of internal radiation by use of cooling panels. Heating. together with expressions for the cooling load requirements. Specially designed reflectors may be used to concentrate the radiation on particular surfaces for removal. so that the resulting energy losses can be estimated in the design of a radiant space heating system. Piping & Air Conditioning. 5. constituting the bulk of filament radiation. Up to 30 percent of the panel energy can be lost through the glass. PANEL HEATING & COOLING 1. Merl. pg. the receiving panel or sink may be operated at a relatively high temperature which may be obtained by use of unchilled water. is extensively absorbed. "Calculations of Direct Energy Losses from Ceiling Mounted Radiant Heating Panels to Fenestrated Areas". No. or the required panel temperature may be computed in accordance with the comfort equation. J. The proportion of energy radiated from a ceiling mounted radiant heating panel directly to a window is calculated. Equations are developed enabling the determination of the effectiveness of enclosed radiant energy sources in heating the room air. From the established relationships. 78.. The results of research conducted by the author show that a maximum of approximately 65 percent of the electrical input to a large incandescent filament may be removed directly from the conditioned space. Because of the high temperature of incandescent filaments. "Effectiveness and Temperature Requirements for Cooling Panels Removing Internal Radiation". The theoretical analysis is supplemented by examples and design charts. Energy Engineering: Journal of Association of Energy Engineers. "Removal of Internal Radiation by Cooling Panels". Internal radiation emitted from electrical lighting filaments may be efficiently removed from a conditioned space by cooling panels. Piping and Air Conditioning. Baker. B-66 . Results are given for 2 x 8 ft and 4 x 8 ft radiant ceiling panels at 175oF located centrally to the window or off-set from the center for various (5 x 6 ft to 8 x 20 ft) window sizes with the panel located horizontally from the window from 0 to 10 ft. Alexander. Vol. Both first (geometric) and second (taking into account the optics of glass) order numerical radiometric calculations are done. June 1952. August-September 1981. both the panel cooling and the convective load components may be calculated for any given panel temperature. 35-48. Merl. Heating. November 1949.G. * 3. for most light colored painted surfaces. ASHRAE Trans. and Supply Penums". these values decrease appreciably ranging from approximately 62 to 48 percent. A further decrease to approximately 52 to 45 percent is accomplished by use of a heat-absorbing panel surface. All radiant panel aspects have been combined and compared to the air distribution component of the total conditioning effect.. On the heating cycle. * * Becker. . The balance in each case flows into the plenum to be distributed by the remaining elements of the floor-ceiling sandwich. depending on the insulation used behind the panel. does the MRT differ from room air temperature by more than 1. The overall radiant panel effects of the floor-ceiling sandwich have been evaluated. For structures not equipped with cooling panels. Boyer. * 4. Vol. Imbedding of the pipes in plaster as compared with exposure of pipes on the back of a panel reduced temperature variation at the surface by 10 deg. the effectiveness of this energy in heating the room air varies with the equivalent overall conductance of the enclosure and ranges approximately from 96 to 68 percent for good and poor insulation. In neither of these two cases. Vertical heat transfer factors for heat removal light troffers acting within a full scale floor-ceiling sandwich system have been determined. about 90% of the conditioning with a supply plenum is due to the radiant environment. The supply plenum condition provides the most favorable radiant environment for occupant comfort. 2. 74. By use of cooling panels possessing a conventional finish. Heating. Temperature differences between various points on the panel surface varied from 26-12 deg. pipes on 7-in.Ceiling Assemblies Incorporating Static. B-67 . with water circulating at 119 F. Only with the supply plenum configuration is the net radiant panel aspect compatible with the conditioning system on both cooling and heating cycles. On the cooling cycle about 15% is directed downward while on the heating cycle about 25% flows into the room. since on the cooling cycle the coolest MRT is obtained and on the heating cycle the warmest MRT is obtained. approximately 65% of the heat from the lighting system is extracted. as indicated by mean radiant temperature. Pt.5 F at the room center. respectively. L. centers. April 1950. Sidney. Return. •* * * * 5. At a flow rate of 70 cfm per unit. 1968.Internal radiation emitted from electric lighting filaments may constitute a major portion of the total cooling load. L. regardless of conditioning cycle or plenum function. "Radiant Panel Effects of Floor . The surface temperatures of a plaster panel containing 3/4 in. "Surface Temperatures of Plaster Ceiling Panels". Piping & Air Conditioning. the heat lag in the coolest part of the panel surface does not exceed 20 to 25 minutes. were obtained at equilibrium conditions for four arrangements of insulation at the back of a panel. W. Human shape factors for approximately 95 percent of the population will agree within plus or minus 7 percent with the data obtained from the dummy which represented an average man. C . R. Effective conductance values are given for the different types of panels for tube spacings of 4 to 12 in. tube spacing and the amplitude of the temperature wave on the panel surface are shown in a series of curves. Howarth. located both above and below the lath. S. "Laboratory Studies of the Thermal Characteristics of Plaster Panels". and Baker. * * * 8. for the majority of people agreement between actual and dummy shape factors will be found to be extremely close. This paper presents the results of laboratory studies on four plaster panels. B-68 . September 1951. V.. L. E.-F. It can be concluded that reasonably accurate predictions of the thermal performance of a wide variety of aluminum ceiling-type heating and cooling panels (including types in which aluminum panels are fastened to tubes by mechanical means such as clips) can be made through the use of the design chart or equations and properly selected values of all contribution factors.6. The panels represented conventional types of ceiling panels. and Koch. Relationships between total panel output. "Optimum Panel Surface Distribution Determined from Human Shape Factors". C. It is shown that the difference between the average tube temperature and the average panel surface temperature can be related to the heat output from the panel by a simple empirical equation. This paper presents the results of an experimental investigation of the shape factor of the clothed human body with respect to energy emitted by floor areas. July 1951. Heating. M. using both non-ferrous tube ferrous pipe. Schutrum. Humphreys. C. Hutchinson. An analytical solution concerning the thermal performance of metal heating and cooling panels is developed and this is reduced to a design chart for easy applications to panels differing markedly from those tested. franks. June 1951. S. M. Merl. F. Heating. The results complement those of an earlier paper in which similar shape factors were reported for the human body (in standing and in sitting position) with respect to energy emitted by wall and ceiling areas. the present study was limited to the standing position. Results of heating and cooling tests conducted on a variety of brazed aluminum ceiling panels are presented fora number of panel and room operating conditions.. Data and curves show the effect of back insulation on panel performance and some information is given on the effect of back plastering. Piping and Air Conditioning. Comparisons drawn between the actual test results and those predicted by the analytical solution indicate good agreement. Piping and Air Conditioning.. Design charts are presented. Huddleston. "Aluminum Ceiling Panels for Heating and Cooling".. Heating. * * * 7. Piping & Air Conditioning. The author points out that designing for panel cooling is not the simple reverse of panel heating. The study was carried out in a college apartment unit under actual living conditions. "Design Factors in Panel and Air Cooling Systems".. May 1955. nor the grey body are adequate for the cooling analysis. Neither the concept of combined surface conductance. "The Mechanism of Heat Transfer Panel Cooling Heat Storage Part II. . The system was designed and operated as a year round unit. Solar Radiation.. Refrigeration Engineering. Charles S.. 1948. Heating. with and without supplemental air supply. it is essential that the methods under comparison shall not produce an end result which will unduly compromise with the production of optimum conditions. Piping and Air Conditioning. C. the author's analyses is extended to the solar load.9. In comparing air conditioning methods. One of the unexpected results was the importance of the radiant cooling effect of the wall and ceiling panels which gave the occupants a comfortable feeling when discomfort might have been expected because of high humidity readings. S. June. Assuming that the air conditioning methods to be compared are capable of attaining the same end result. The author has attempted to present the theory of panel and conventional air cooling systems and to indicate possible courses of panel cooling design. C. S. and the economics be determined not solely on the owning and operating cost of the air conditioning but on the owning and operating cost of the entire building. * * Leopold. and for two types of ceiling finish. * * Leopold. July 1947. The cooling of the water was done with an evaporative cooler. R. Phenomena of heat storage and panel cooling are discussed in terms of the mechanism of energy transfer source to atmosphere and enclosure. Data obtained and presented in part in this paper seem to indicate that panel cooling can be used successfully to provide summer comfort in the residence tested. Test data are presented for a continuous cooled ceiling for various types and sizes of luminaires. Piping & Air Conditioning. but this paper deals only with the cooling phase. In this paper. Irwin. Heating. * 12. May. "The Mechanism of Heat Transfer Panel Cooling Heat Storage". Test B-69 . 1951. R. * * Leopold. Refrigerating Engineering. * 10. "Panel Cooling for a Residence". The air conditioning design should be related to all elements of building construction and use. the selection of a particular form of air conditioning is a matter of economics. * 11. 58. Vol. F.. Clarence A. * 15. J... * 13. A four-room one-floor. Mostly descriptive in nature. Points out the benefits of reducing air motion and the bringing in of outside air. Refrigerating Engineering. Comments are made on comfort reaction of occupants and on economic justification of expenditure for reduction of heat loss of the structure. * * Mills. "Performance of an Electrical System of Panel Heating with Four Stages of Insulation"..data are presented for the performance of panels and the overall heat balance for an enclosure with a continuous cooled celling. mineral wool insulation to floor. This article reports on the design and first test results of cooling and heating by radiant cooling and heating panels on the ceiling. R. Refrigerating Engineering. January.. 1955. Jan. It presents a description of a single residence set up for radiant cooling and heating. A discussion of the benefits of radiant heating and cooling panels for residential applications. * * * 16. Mills. A. No. "Sensible vs Latent Heat Removal in Radiant Cooling". Piping & Air Conditioning. * 14. (2) adding 2 in. t A discussion of radiant cooling in industrial and commercial applications where high latent loads are present. Operating results and cost data are reported for the original construction and for three conditions in which the heat loss was progressively reduced by the following steps: (1) change to double glazing of windows. was operated through the 1947-48 heating season. C. Refrigerating Engineering. B-70 . and Schreiber. Several basic solar load properties which are now incorporated into the ASHRAE design procedure are brought out here. one window. * * Mills. A. C. and (3) addition of 2 in. completely heated by panels of electrically conductive rubber. of mineral wool insulation to ceiling. occupied residence. 11. Heating. March 1958. "Reflective Radiant Conditioning Can Provide More Comfort at Less Cost". "Residential Cooling by Reflective Radiation". November. * * Lorenzi. J. and various combinations of glass and shading devices. 1949. 1950. H. A floor slab on the earth is the intended panel heating application. * 18. Effective field of application of radiant heating systems with suspended radiating panels taking account of the shape of the premises. Vol. That study showed that energy use could be reduced by as much as 40% using radiant cooling panels. with a solid conducting medium adjacent to this surface. L. A general mathematical solution was obtained and electrolytic analogue was employed for establishing application data because of the complexity of the mathematical treatment. Many claims are made but not a great deal of data is given. "Year-round Residential Conditioning By Reflective Radiation". Results are presented in the form of the thermal resistance between the pipes and the opposite isothermal slab surface. V. . 6. T. value of infiltration heat exchange. B-71 .. Shilkrot. November 1950.Heating.. 93. moisture content of the premises. * * * 19'.63). the authors reported results of research on the effect radiant cooling panels had on stratification in high bay buildings. A. Hulbert. In 1982. Clarence A. 1.. Mills. * 20. 1987. C. Theoretical and experimental research works being conducted in the Soviet Union gave opportunities to elaborate an effective structure of radiating panels (convective heat transfer from 1 kg of metal equal to 65 watts. Franks. Refrigeration Engineering. B. Clima 2000 . Pt... W S Kongres . a share of radiant heat transfer is about 0. Nottage. "A Computer Program for Radiant Cooling of High Bay Buildings". ASHRAE Trans. J. "Heating of Industrial Buildings with the Help of Suspended Radiating Panels". 0. and Singh. and under conditions such that the opposite slab surface may be taken as isothermal.17. E. Schutrum.. Solovyov. and a scheme of their location in the premises. "Heat Flow Analysis in Panel Heating or Cooling Sections". * * Olivieri. Ventilating and Air Conditioning Systems.. Heating. B. 1985. L. F. A method is suggested for the design of radiant heating systems which make possible the determination of the necessary amount of radiating panels.W S Messe... has been presented in the paper. J. Radiant heating ensures a high level of comfort and low heat consumption. Piping and Air Conditioning. E. This article is a discussion of radiant cooling and heating for residential applications. The heat flow has been studied for the case of a row of pipes or tubes imbedded in a slab and tangent to one surface. A. L. May 1953. * * Naumov. V. This seems to be due principally to the fact that the convection conductances obtained for the.In this latest study.S. 22. L. C. The calculation method used finite differences to predict the amount of heat transferred through the roof and into the floor slab. With no infiltration the AST minus the room air temperature was directly proportional to the ceiling temperature minus the AUST. ceiling panel appear to be much lower than presently published data would indicate. Contains only descriptive information. For constant values of ceiling temperature and AUST. Heating and Refrigeration News. Parmelee. The total heat output of the ceiling panel was much lower than is given in some presently used design methods. As defined before. 3. 1985. Air Conditioning.. 1952. a definite relationship was established between room surface temperatures. 2. The heat flow due to radiant exchange between a given surface and the rest of the room may be opposite in direction to the heat flow due to the convective exchange between the surface and the ambient air. Humphreys. B-72 . and the temperature of the room air at the 60 in. For this room. V. A description of a radiant panel cooling system that will be marketed in the U. Piping and Air Conditioning.S.". the difference between the AST and the room air temperature increased as the rate of infiltration increased and as the temperature of the infiltration air decreased.. M. infiltration air rate and temperature. 21. December. G. Schutrum. "Heat Exchanges in a Ceiling Panel Heated Room". . and the AUST is the area-weighted average temperature of the unheated surfaces of the room. . a computerized calculation method suitable for use on a personal computer is presented. "Radiant Cooling Panel will Get Tryout in U. level. F. the AST is the area-weighted average temperature of all the room surfaces. For this room and the specific test conditions reported here the following observations can be made: 1. Heating. October 21. 4. The surface temperatures of neutral walls were not necessarily the same as the temperature of the ambient air but were dependent upon the heat balance between the radiative and convective heat exchanges. 5. The performance of a panel heating system in a space having a non-uniform surface temperature environment can be predicted with satisfactory accuracy on the basis of the area weighted average unheated surface temperature (AUST) of the space. In general. 4. panel surface temperatures may be used for calculation of the upward heat flow. centers with back-plastering. Piping and Air Conditioning. . "Further Studies of the Thermal Characteristics of Plaster Panels". C M . The heat transfer within a plaster panel is related to the heat output from the lower and upper surfaces of the panel and can be expressed in terms of an effective conductance. However. these tests were made under such a variety of conditions that it seems reasonable to assume that they may be applied with satisfactory accuracy to any ordinary structure. Insulation on the back of plaster panels serves the following purposes: (a) reduces upward heat flow: (b) increases the average •panel surface temperature for a given tube temperature or conversely: (c) for a given downward heat flow. 1. For all practical purposes. June 1953. the net effect of these trends may be considered negligible for design purposes. February. C. M. and Humphreys. L.23. Schutrum. . F. It has been demonstrated that for panels having tubes above metal lath. F. 5. 1954. good tube embedment and good contact between tubes and lath are prerequisite to good heat transfer. Piping and Air Conditioning. B-73 . Heating. the heat transfer of a panel with tubes on 6-in. However. 3. 2. L. Heating. and Humphreys. Furnishings in a panel heated space tend to reduce the heat output of the panel. 24. "Effects of Non-Uniformity and Furnishings on Panel Heating Performance". Schutrum. The authors' conclusions are as follows: 1. permits operation with a lower tube temperature than would be required if the panel were not insulated. The addition of back-plastering to panels constructed with tubes above metal lath increases the heat transfer to the panel surface. Tests from which the following conclusions are drawn were madein the Environment Laboratory. centers without back-plastering is equivalent to that of a panel with tubes on 8-in. 2. and increase the room air temperature. This is defined as the downward heat flow in Btu per (hour) (square foot) divided by the difference between the average tube temperature and the average panel surface temperature in Fahrenheit degrees. and Vouris. D. 2. * * Schutrum. C . 3 and 4. the authors arrived at the following conclusions: 1. L. Ceiling panel cooling is an inversion of floor-panel heating. "Effects of Room Size and NonUniformity Non-Uniformity of Panel Temperature on Panel Performance". January. In a typical test. John.3. convective. 3. surface temperature of which is above normal inside dew-point temperatures.. Cleveland. F. Coverings for heated floor panels should be selected to provide a minimum of resistance to heat flow from the panel to the space. The maximum test temperatures of the heated portions of floor and ceiling panels were 95 F and 140 F respectively. Infiltration of warm air and internal. * 26. B-74 . September. and the performance of a cool ceiling panel can be predicted from the performance of a warm floor panel. the surface temperature of that panel and the surface temperature of the heating medium must be considerably increased to maintain the same heat output to the space that would be obtained from the bare panel. The preliminary conclusions from this study were: 1. Vouris. and Min. Within the range of conditions tested. F. * 25. heat sources must be considered for design purposes. From laboratory tests. a ceiling at 10 deg F below room temperature absorbed 20 Btu per (hr) (sq ft) of ceiling area. may safely be used without correction for the design of panel heating systems for any space of normal size and proportion. Pining and Air Conditioning. L. ceiling or floor panels comprised of heated and unheated sections have the same total heat output and produce the same room air temperature as if the entire area were heated to a uniform temperature equal to the area weighted average of the heated and unheated surfaces. .sq ft). 1954. 4. the amount of this temperature increase was found to vary from 27 deg to 60 deg for the various combinations of carpets and pads tested. 2. Heating. When a floor covering is laid over a heated floor panel. T. The effects of room" size on the performance of floor and ceiling panels are relatively small so that the heat transfer relationships developed in the standard room and reported in References 2. "Preliminary Studies of Heat Removal By a Cooled Ceiling Panel". Piping & Air Conditioning. For a heat output of 25 Btu per (hr) (. J. An appreciable amount of sensible heat can be removed from a room by a cooled ceiling panel. Heating. Ohio. * * Schutrum. 1955. C . in part. variations in room-air temperatures were found to be small. Schutrum. L. Room furnishings can be neglected as they affect the heat pickup by only approximately 5 percent. F. T. 1956. 1957. 8. and 58 percent was transferred to the room air by convection. For indirect lighting. 28. (c) the radiation from the lights to the panel can be added to the panel pickup as determined (a). 2. . and Min. Mean radiant temperatures were in general higher than the room-air temperature. For the particular fixtures used in this study. 3. August. Cleveland. 5. Room-air temperatures. and Min. "Cold Wall Effects in a Ceiling Panel Heated Room". In a test room heated by a ceiling panel. Heating. these values were 10 percent visible radiation. the air movement. Cleveland. and 69 percent convection. Ohio. by the conditioned air. B-75 . F. and the division may be determined by Fig. In the living space. air velocities. # The room-side temperature of a wall or all-glass wall exposed to winter outdoor conditions affects the air temperature. 27. wide and the panel temperatures selected to maintain a constant 70 F room-air temperature. The ceiling was heated in panels 4. The authors reached the following conclusions. Non-uniform wall and floor temperatures can be represented by area-weighted average temperatures (AUST) in calculating heat pickup by a cooled ceiling. T. approximately 10 percent of the energy supplied to the direct lighting system was radiated in the visible wavelengths. Heating. "Lighting and Cooled Air'Effects and Panel Cooling". The heat pickup by a cooled ceiling-panel and conditioned air may be summarized as follows: (a) The sum of normal heat gain through the room surfaces plus the radiation from lights which falls on the walls and floor is removed in part by the panel and. and 12 ft. and room air velocities were neglibible except near the floor. L. Piping & Air Conditioning. 21 percent long-wave radiation. Schutrum. Piping & Air Conditioning.4. and the radiant conditions within the room. C . 32 percent was emitted as long-wave radiation. (b) The convection heat gain from the lights is removed almost entirely by the conditioned air system and may be added to the convected load as determined in (a). Ohio. one whole wall was cooled to simulate the inside temperature of an exposed wall. 1. November. and mean radiant temperatures were measured under steady-state conditions and are reported. or by Equation 3 of the paper. . B. 1982. . 1. 2. for indirect lighting. 2. Pt. B. * 29. The conclusions from this study were the following: 1.. The position of the panel has no significant effect either on the temperature stratification or the energy consumption. "Effect of Radiant Cooling Panels on Temperature Stratification". and Olivieri. V. Final Report. T. 3. and Olivieri. 4. 4. The use of radiant cooling panels produces a slightly greater degree of temperature stratification as compared to those produced by air systems at all loads.3. * 30. 88. J. either on the temperature stratification or the energy consumption. The cooling load factor Fc decreases as the lighting load increases. The position of. This includes the heat radiated directly to the ceiling. compared to that for the most efficient air system (using combined supply and return air system). ASHRAE Trans. The conclusions of this ASHRAE project were listed as follows. ASHRAE. The energy consumption for radiant cooling system is 8 percent less. J. . B-76 . these changed to 18 and 12 percent respectively. "Effect of Radiant Cooling Panels on Temperature Stratification under RP-260". and 24 percent to the walls and floor. 1981. * * Singh. Approximately 16 percent of the energy input to the direct fluorescent lighting fixtures was radiated to the cooled ceiling panel. More than 20 percent of the energy supplied to the lighting system can be removed by a cooled ceiling panel. the panel has no significant effect. 3. and that which is reradiated from the other room surfaces. August. The use of radiant cooling panels produces a slightly greater degree of temperature stratification when compared to those produced by air systems at all loads. T. The energy consumption for radiant cooling system is 8% less as compared to that for the most efficient air system (using combined supply and return air system). The cooling load factor Fc decreases as the lighting load increases. 2. 4. * * Singh. With its simplicity. 1955. Staff Members of the ASHRAE Research Laboratory. Subcommittee of TAC. Heating. B-77 . it retains engineering accuracy appropriate to the usual applications of panel heating in residential and commercial buildings.31. Piping & Air Conditioning. "Thermal Design of Warm Water Ceiling Panels". The design procedure here presented is based on research and provides a reliable means for the thermal design of ceiling-type heating panels using warm water as the heating medium. December. 92.. Humphreys. ASHRAE Trans. This work has been reported in the 8 research papers listed in the Bibliography at the end of this paper. March 1952. which simulated various conditions of construction and outdoor temperature. The ceiling height was also adjustable and ventilation was variable. A convective heat transfer computer model was utilized to predict energy consumption of a radiant panel heating and cooling system. * 33. The simplified procedure provides a panel design to maintain the desired room air temperature for the selected outdoor conditions. "A Numerical Study of Heat Transfer in a Hydronic Radiant Ceiling Panel".. Findings from these studies have been weighed as to importance. a system load adjustment should be considered for the radiant panel model. and Parmelee. 1986. . 24. Z. and significant insights into B-78 . D. The room could also be divided into smaller segments. The panel system responded faster than convective systems and maintained a more uniform mean radiant temperature in the room. Pt. and the design procedure is restricted to situations in which the area weighted average temperature of the walls. and a combination of radiation and convection heat transfer. the radiant panel heating and cooling system can be justifiable on a life cycle cost basis. M. Heating. 62. located in the ASHVE Research Laboratory. adiabatic. and glass does not differ greatly from room air temperature. Finally. E. "The ASHVE Environmental Laboratory". Several different numberical solution schemes were investigated.. 1. showed that this near-equality of the two temperatures normally prevails.The basis of this procedure is a body of experimental data obtained at the ASHRAE Research Laboratory in a comprehensive program planned and guided-by Technical Advisory Committee on Panel Heating and Cooling.. and Pate. P. * * Tasker. * * Zhang. New York. G. * 32. ASME. the floors. Room air temperature is the selected criterion of comfort. . V. Vol. Therefore. Numberical Methods in Heat Heat Transfer. HTD-Vol. Vol. provides facilities for the study of panel heating performance and also for further study of human reactions to environment involving various combinations of temperatures of air and interior wall surfaces. CM. Piping and Air Conditioning. * * Weida. Three different types of boundary conditions were required: isothermal. and then trimmed to meet the needs of a simple but accurate design procedure. C. E. Temperatures of all surfaces or portions of them may be varied at will to simulate field operating conditions. The new Environment Laboratory. "Life-Cycle Cost Analysis of Hydronic Radiant Panel". 1986. The heat diffusion equation was used to model numerically a radiantceiling panel for both steady-state and transient heat transfer. The room-scale tests. * 34. B. . the computation time was less for the explicit method as compared to the implicit method because of the radiation boundary conditions. including tube spacing. plaster thickness. and convection heat transfer rate. The explicit method was found to be the most effective method for solving both the unsteady-state heat diffusion equation and the steady-state Laplace equation.the advantages and disadvantages of each scheme were obtained by comparing the results. The output of the transient model was the temperature history of the radiant panel ceiling including the final steady-state temperature distribution and the heat flux from the panel. Several design considerations are investigated using the numerical model. B-79 . The heat transfer characteristics of the heating panel as predicted by the transient numerical model are also discussed herein. Specifically. B-80 . Vol. 1978.1984 Systems Handbook). and system precautions. * 2.Chap. Refrigerating and Air Conditioning Engineers. system efficiency. 1983. avoidance of temperature stratification and air motion there is a considerable saving (20%) in energy. Refrigerating and Air Conditioning Engineers. " • * * * 3. "High I n t e n s i t y Infrared Heaters .. Tables are given for gas infrared unit efficiencies. complete building heating. This information was prepared by the ASHRAE Task Group on Radiant Space Heating. The current chapter in the ASHRAE Handbook describing high intensity infrared radiant heating. April. 18". High temperature radiant heating in large spaces offers physiological advantages. "High-Intensity Infrared Heaters". 30 in the 1983 Equipment Handbook) and Infrared Radiant Heating (Chap. American Society o f Heating. Atlanta. "High Intensity Infrared Radiant Heating . Gewea GmbH and Co. 1984 Systems Handbook. ASHRAE Journal.Chap. GA. The current chapter i n the ASHRAE Handbook describing infrared h e a t e r s . 4. Atlanta. 1984. system design principles. * * Baumanns. 18 . Warme Gas International. Cool air and comfortable radiation reduce fatigue and the amount of particulates in the air. GA. American Society of Heating. and is the basis of the information presented in the ASHRAE Chapters on Infrared Heaters (Chap. gas infrared energy generators.H. electric infrared energy generators. ASHRAE Task Group. INFRARED HEATING 1. No. ASHRAE. It considers such items as applications. * 4. "Gas IR Heaters for Heating Large Spaces". Because of the low temperature of the air. * * ASHRAE. wind or draft effects. H. controls. (In German). 1983 Equipment Handbook. 27. Figures are given for determining the heat requirements for industrial buildings. December 1963. spot heating. Monchengladbach. 30". Piping and Air Conditioning. D. "Spatial Radiation Patterns for Infra-Red Heaters". dress. Belsey. R. nomenclature. and Benseman.. (2) Reliable controls permit balancing output with the heating requirements of the moment. The assumptions used to simplify the computation do not introduce significant errors and the method produces data of adequate accuracy for all practical purposes. "Application and Selection of Electric Infrared Comfort Heters". (3) There are no air contaminants produced by the heating system. * * Boyd. 3. and engineering data and standards are needed sorely in this field.5. recommended design procedures. ASHRAE Journal. This article presents a comprehensive review of what is known about infrared comfort heating by quartz lamps and tubes and metal sheathed heaters. 1969. An example case is given. Heating.. and others. those circumstances of position. B-81 . F. Many places where no practical means can provide real comfort can be made tolerable by infrared. This article points out that any space can be heated by infrared to improve comfort. Building Science. Where any other means is practical. 133. Where infrared is indicated for comfort heating* electric infrared is advantageous because: (1) elements approach point or line sources. • * * Boyd. Authoritative definitions. Nov. A method of determining "watts input density" based on the amount of radiation delivered to a surface is also suggested. wind conditions. Information such as this will assist designers of high intensity infra-red heating installations. When evaluating high intensity infra-red heaters.. infrared will not be the preferred method in most cases. R. 1960. allowing excellent control of pattern (keeping radiation where it is wanted) . Many places impractical to heat any other way can be heated effectively and economically by infrared. Robert. October 1962. the data from a limited number of radiation measurements can be used to compute spatial radiation patterns for a range of mounting heights and tilts. G. * 7. Vol. p. "What Do We Know About Infrared Comfort Heating?". Other means should be investigated thoroughly before deciding to use infrared. which cannot be so accurately measured but which can seriously affect the adequate operation of an infrared installation. and perhaps more importantly what is thought to be known i. L.e. * 6. A emthod for evaluating the radiation intensity distribution is presented. movement. "Gas-Fired Radiant Heat". Boyd. other details and characteristics of elements. 87. * * Bryan. In these unrealistic extremes the MRT for 100% radiant heat input was 8 deg C greater than 100% convection heat input with the result that the space air temperature required for equal comfort is reduced 8 deg C. Electrical Engineering. Pt. the heat loss from the building could be supplied by convection heat input directly to the space air or by radiation transfer directly to the internal surfaces of the building. The inside surface temperatures of the space with the exception of the floor remained below space air temperature for both radiant and convection heat input. This one-to-one reduction results from an increase in radiant coefficient for the clothing surface of 4%. these surface temperature increases result in a heat loss greater than had been expected. . Thus results could be obtained for the model for the extreme conditions of 100% radiant to 100% convection heat input.8. This was primarily an analytical study to determine the fundamental heat transfer characteristics of a radiant heating system. Although there is a shortage of available data. V. Mech. 87. "Control of Electric Infrared Energy Distribution". * * Bryan. no room air temperature gradient. L. Control of patterns of electric infrared radiation offers opportunity for material improvement in effectiveness for many applications. B-82 . references given indicate that the nature and efficiency of the elements producing the infrared are not nearly as important in effectiveness of the system as is control of the pattern of the radiation produced. no radiant energy to occupants) make the results questionable. W. A comment on the validity of data in an article in Mechanical Engineering and AGA . . Feb.. Vol. R. ASHRAE Trans. At the same time. Some technical discussion of control of radiation patterns is given. L. and their application have been neglected or overlooked. * 9. Engr. . Several assumptions (high surface emissivity. March 1965. W.. L. The growth in popularity of infrared heating systems for personnel has been accompanied by overemphasis of some details and characteristics of specific elements. He objects to the use of GIR factor for rating radiant gas burners. 1981. Intrinsic to the analytical model.92.Bulletin . 1963. For radiant input. * 10. 1. "Comparative Energy Requirements of Radiant Space Heating". p. 103. fixtures and systems. The only electric infra-red generator which had a value this high was the quartz lamp. * * Cohn. However. 8.4 to 16. No. (40-70%). Co.. A. which is essentially a burner emissivity factor. 1977. Lisa.0 microns. Vol. November 1982. powered and catalytic gas-fired infra-red generators. American Gas Association. 155 +. * 12. Gas-fired burner total normal radiation values as high as 62. Medium-intensity radiant heating provides a cost-effective. All of the comparisons were made with steam radiators which can be "energy inefficient" if not properly maintained. Pg. N. Buckley. Converting energy savings to cost savings demonstrates the cost advantage that is-available with radiant heating. This GIR factor takes flue gas radiation into account since it adds considerably to incandescent gas burner surface radiation. 92. "Gas-Fired Medium . * 13. * * . 1. Vol. 4.11.Intensity Radiant Heating Provides a Cost-Effective. P. 1986. "Radiant Heating Units Net Big Savings in Special Cases". Energy User News. A news article concerning the advantages and disadvantages of gas infrared heaters. April. Research Bulletin No. energy-efficient alternative for space heating. * 14. most of this lamp's energy was short wavelength energy (about '65 percent below 2 microns).. Plant Engineering. 31. and Seel T. and those of different types of electric generators are developed and tabulated. "Conserving Energy with Infrared Heating". Efficient Space Conditioning Alternative". 7.. Radiation characteristics of atmospheric. Tube-type infrared medium-intensity radiant heating systems installed in four diverse applications demonstrated substantial energy savings. D.. 1962. pt. Focuses most on the economics of the systems and where they are most suited. Nov. ASHRAE Trans. No.E. No. "A Study of Infra-Red Energy Generated by Radiant Gas Burners". . Technical Pub. 92. B-83 . 47. Vol. These data show how much total normal energy is emitted by the different types of radiant heat sources at different temperatures and the spectral distribution of this energy in the infra-red spectrum from 1. A Gas Infra-Red Radiation (GIR) factor was developed for gas burners.800 Btu per hour. * * DeWerth. A general discussion of infrared heating* systems and how they affect comfort. W. square foot of burner surface were measured. Tha author describes the various types of gas infrared heaters. "Radiant Heating". 26-28. * * Diamant. Brief data are presented to show the effect of special burner coatings on the burner total normal radiation and red brightness temperature. W. The Heating and Air Conditioning Journal. "An evaluation of the available literature pertaining to I-R energy production by both gas and competitive means is presented. Ohio. No. R. (3) work on problems of current applications and develop recommended new applications of gas-fired I-R burners.76 as the tube or surface temperature varied from 150oC to 500oC. Cleveland. Troop Publ. * * DeWerth. the porous refractory burner. This article compares the radiant and convective heat output of various English infrared radiant heaters. and compared to I-R energy generated by competitive means. D.900 F) I-R burners". ASHRAE Journal. * 16. * * DeWerth. "Gas-Fired Radiant Heat". "Literature Review of Infrared Energy Produced with Gas Burners". but its total normal radiation was somewhat lower than that measured for comparable gas equipment. Gas-fired I-R burners are described and evaluated. 50. American Gas Association Laboratories. The necessity for future research is indicated. 86. These include the radiant tube. and (4) develop information on low temperature (below 750 F) and high temperature (above 2. Research Bulletin 83. W. They are: (1) measure the emittance spectra of gasfired I-R burners. * 17.The only electric infra-red generator which emitted energy qualitatively equal to that of the gas-fired generators was the quartz tube. Vol. 1964. The abstract of this work done in 1960 is reproduced here. pg. M. Four subjects for future work are recommended. Mechanical Engineering. which have been developed since 1950. Vol. Additional details on the performance B-84 . 577... the direct fired refractory burner and the catalytic burner.. They indicate that the ratio of radiant to convective heat output varied from 0. 1980. February. D.63 to 1. (2) measure the absorption spectra of possible loads not covered by the literature. E. The literature search has indicated that the use of gas-fired I-R burners should become a field of ever increasing gas usage. * 15. 1960. . Nov. The burner characteristics are presented including the spectral radiation curves and their typical applications.. Many current applications of I-R are described and an extensive list of possible future applications is presented. electric and gas-fired infrared space heating systems in order to determine thermal comfort level.. "Field Evaluation of High Intensity Infrared Space Heating Systems . No. A good discussion of the placement of radiant heaters is given. energy consumption and cost are presented. • * * Heath. The author discusses three types of direct-fired radiant heating systems: (1) gas-fired infra-red panels. and the need for evaluating condensation. The author discusses heat loss (reduced by 80-85%) energy requirements (reduced by 80-85%). Vol. A. * 19. Automation. . October 1972. 571. * 21.of these heaters is also given. 6-9. 49. * 20. Data for the study came largely from actual field study and analysis. * 18. 1967. operation. The information is just general in nature. There were no general summaries presented. * * Field. B-85 . and (3) air-heated radiant tube. ASHRAE Journal. * * Faucett. Heating and Ventilating Engineer. "Controllable and Efficient Infrared Radiant Heating". February 1975. energy consumption.. W. and installation and operating details. This study program was designed to validate the conclusions of RP-41 by measuring in the field the performance of high intensity. Technitrade Journals Ltd. Data is presented but in the form in which it exists. unit placement. pg. * * Hunter. it is difficult to evaluate. controls. . George. Robert K. Final Report. ASHRAE."Direct Fired Radiant Heating Systems". The author presents descriptions of electric infra-red heating systems and their applications. Factors such as efficiency.Research Project RP-98". applications. J. A general discussion on the use of gas-fired infrared heater is given. ventilation requirements. (2) gas-fired radiant pipes. "Basic Infrared Heating Applications". Dec. A. 1972. ASHRAE Trans. buildings were poorly insulated and subject to drafty conditions when there was a wind. 3.5 met. Faucett of York Research. Infrared heating involves not only the air temperature in the space but gains due to the direct heating effect of the radiation as well. June 1968. 23. 1. "An Economic Study of an Electric Infrared Space Heating Installation". Pt. The data were widely scattered due to conditions beyond the control of this test.. this probably was not sustained for any length of time. the energy is used more effectively and efficiently because the heat is more confined to the lower level where it is needed. With infrared heating. Radiant heat appears to offer distinct advantages. Intermittent operation tends to lead to complaints of dampness and cold floors. In a conventional system the reverse would have been true. Vol. However. Comfort is compromised when systems are manually operated or operated on an -intermittent basis. The installed heating capacity for electric infrared can be as much as 15% less than calculated heat loss for other types of heating systems. care must be exercised in designing the system to assure uniform distribution of radiation to the occupied areas. These areas should be protected from excessive drafts caused by large open access doors if possible. Janssen. It is doubtful that the steady state activity exceeded about 2. The data obtained actually showed up to one-third less than the calculated heat loss would indicate. It appeared that the estimates of physical activity were too high. John E.. The operative temperature was an average of 5. A. This paper is an analysis of the data collected in RP-98 by J. especially in industrial installations. 2.. It is estimated that this resulted in at least 20% reduction in fuel over convective heating systems. Infrared heating has a lower roof heat loss because there is less stratification. With exception of the swimming pool.9 deg C) higher than the dry-bulb temperature. This result would not be peculiar to radiant systems but could be expected with convective systems also. Even so. "Field Evaluation High Temperature Infrared Space Heating Systems".5 to 4. Therefore. people tended to indicate general comfort. Morelli. In fact. However. B-86 . 1976. W. it seems that an installed capacity less than 80 to 85% of calculated heat loss at the design outdoor temperature is risky at this time.2 deg F (2.5 met. tests showed temperatures at the ceiling several degrees lower than at waist level. This form of heating for commercial and industrial type buildings with high ceilings has these advantages: 1. 82. in the same area there can be in effect two or more comfort temperatures called operative temperatures by installing different densities at different locations in the space. Although certain tasks may have required 3.22. J. ASHRAE Journal. c. b. DC. Exhaust openings for removing flue products shall be above the level of the heaters. .3 Listed heaters shall be installed with clearances from combustible material in accordance with their listing and the manufacturer's instructions. Washington. In locations used for the storage of combustible materials.18. Unlisted heaters shall be installed in accordance with clearances from combustible material acceptable to the authority having jurisdiction. National Fuel Gas Code. This manual is the third edition of NEMA Standards Publication HE 3. 1983. b. 6. Heaters subject to vibration shall be provided with vibration isolating hangers.1. 1984.10 and 5. natural or mechanical means shall be provided to supply and exhaust at least 4 cfm per 1.18 Infrared Heaters. Infrared Application Manual. 1984. ANSI Z 223. Where unvented infrared heaters are used. HE 3-1983. 6. signs shall be posted to specify the maximum permissible stacking height to maintain required clearances from the heater to the combustibles.1. Combustion and Ventilation Air: a. first issued in 1971 and updated in 1976. This code listed a short section concerning infrared heaters which is given below.2 Clearance: a. 25.18. American Gas Association.1 Support: Suspended type infrared heaters shall be safely and adequately fixed in position independent of gas and electric supply lines.11.1. National Electrical Manufacturers Association. Hangers and brackets shall be of noncombustible material. 6.4 Installation in Commercial Garages and Aircraft Hangers: Overhead heaters installed in garages for more than 3 motor vehicles or in aircraft hangars shall be of a listed type and shall be installed in accordance with 5.24.18. 6.18. Standard Publication No. NEMA. HE 3-1983 addresses the infrared heater both as a unit and as part of an overall system involving other factors affecting comfort B-87 . It reflects much of the current technology involved in infrared heating for improvement of comfort level. As NEMA Authorized Engineering Information. 6. and its purpose is to help potential users determine how infrared electric heating can be a viable method of meeting their space conditioning needs.000 Btu per hour input of installed heaters. 14-18. June 1968. compensating for both the ambient air and radiant temperature effect upon occupant comfort. B-88 . R. Emphasis was placed on radiant systems that are currently commercially available. There is a large potential field of expansion in the use of gas for commercial and industrial space heating. H. demonstrates substantial operating economies for the infrared system. R. and how they compare with each other. which are currently available. May/June. * * Pam. C. Building Materials. Comparison of the electric infrared energy consumption with the calculated consumption of a conventional system. Vol. "Five Years Operation of an Industrial Infrared Heating System". * 27. A discussion about the types of radiant heating systems. Alzeta Report No. 630. Burner Survey for a High Efficiency Gas-Fired Heating Unit..level such as draft and humidity. 1973. Troup Publ. P. Len. 21. and Kesselring. ASHRAE Journal. January 1984. The various applications of infrared heaters * 26. The Heating and Air Conditioning Journal. The authors' conclusions were stated as follows. M. * 29. particularly when the structure U value is high. * * Taylor. July-August 1984.. 54. Nine different burner types were reviewed and evaluated against criteria established by the hydronic heating unit manufacturer. * * Sanford.. 18. are described as well. 1. Teledyne Laars. * * Simmons.. J. A survey was undertaken to evaluate the suitability of a variety of burner types for use in a new concept. high efficiency residential hydronic heating unit. F. 84-706-104. Five years of operating experience with this installation has demonstrated the adequacy of the infrared system design for maintaining the design operational temperature at 50 F. A good discussion of radiant vs warm air as well as good points brought out about various types of radiant systems. Results of the survey show that a porous fiber matrix burner can most easily be incorporated into the heating unit design and meet the operational criteria. Pg. *• 28.. No. "Radiant Space Heating". "No Problem with Radiation". L. savings come from reduced infiltration losses. M... "Development of a Standard Test Method for Measurement of the Radiant Heat Output of Gas-Fired Infrared Heaters". and Macriss. B-89 . V. T. and the floor beneath the radiant tube and not by the walls since they are cooler.6-1982) method for determining total infrared heat output from commercial gas-fired infrared-tube space heaters is not sufficiently broad to encompass all of the equipment currently in the marketplace. the reflector assembly causes most of the energy to radiate in a downward direction and. The grid size would be simply too large for practical use. R. ASHRAE Trans. Second. No. Finally. long straight or U-shaped tubular heaters cannot be evaluated under the spherical grid approach specified in the standard. "The current American National Standards Institute (ANSI) standard (Z83. This difference is largest in the lower 4. 4. The first conclusion to be drawn from this study is that steady state for the tube is reached in approximately 12 minutes. 1. 5. . This can be important for comfort after a large decrease in the temperature of the inside air and inside surfaces. Third. Overhead radiant systems must have effective insulation or reflectors such that the radiant heat component is paramount. 1. By the use of radiant heating and localised burners. S. V. . since convected heat rises to the roof space where a substantial proportion is wasted. For example.. and lower inside wall temperatures. the variation in wall and floor temperatures depends on the orientation to the tube. ASHRAE Trans. The utilisation of gas in the most efficient types of equipment will play a significant part in the reduction of atmospheric pollution as well as making a contribution to the conservation of national energy supplies. therefore. 93. A. R. "An Experimental Study of an Installed High Temperature Radiant Heater and Enclosure". Vol. R. Pate. the temperature of the air is lower in the garage area heated with the radiant heater than in the warehouse area heated with a forced-air heating system. 1987. the air is heated by the radiant tube. . 1. Pt. >v * Zawacki. Fourth. gas for heating at a premium price can compete effectively with oil.2. the wall temperatures vary at levels below 3 m (10 ft) and the floor temperatures are highest directly below the radiant tube. Significant changes in the floor and wall temperatures may affect comfort.5 m (15 ft). 92. * * * Trewin. 1986. For example.. Heat loss through the ceiling was not reduced significantly. the reflector. The concept of environmental temperature is accepted as a satisfactory measure of comfort temperature taking due regard to both air temperature and mean radiant temperature. and Nelson. * 31. less air movement on the inside walls. during the on-off cycle. Pt. Huang. The procedure was tested in the laboratory and found to perform satisfactorily. radiometer window material. The measuring procedure developed includes several parameters that were carefully assessed and defined. Current test procedures were evaluated along with instrumentation used to measure infrared heat rates." B-90 . calibration techniques.000 Btu/h gross input. the need for radiometer water cooling.An approach to extending the procedure to other heaters has been developed through a literature survey and a laboratory evaluation. A total radiant output of 40.500 Btu/h was measured emanating from a straight tube heater of 104. and measurement of background radiation. specifically. A procedure was developed for measuring total radiant output based on the use of a 180° view angle radiometer and a cylindrical grid surrounding the heater. N. baseboard. increasing comfort and improving temperature recovery capability. NY. The Industial Press. Adlam. * 2. Discusses types of systems (ceiling. 1947. Radiant Heating. "Engineering Principles Support an Adjustment Factor When Sizing Gas-Fired Low-Intensity Infrared Equipment". . 1984. 1.. Much of the information is based on practical experience and some measurements that are given are limited to specific cases. * •3. therefore.I. The conclusions from this study as given by the author are as follow. 1987. wall. . 1984 Systems Handbook. N. * * Buckley.Chap. Presents layout details for heating and cooling as well as snow melting. Specific design steps are given but need to be updated with data accumulated over the last forty years. 8". T. + Reduced air temperature stratification with low-intensity gas-fired radiant heating systems reduced structural heat loss significantly. floor. * * ASHRAE. DESIGN PROCEDURES 1. "The demonstrated performance characteristics of low-intensity radia. V. GA. "Panel Heating and Cooling Systems . 93. One of the original design texts for radiant heating systems. and electrical) panels. Factors that contribute to this conclusion are: + Floor temperatures are elevated above ambient by low-intensity gasfired radiant heating systems without heating the air. Pt. Refrigerating and Air Conditioning Engineers. The current chapter in the ASHRAE Handbook describes panel heating and cooling systems. + Improved mean radiant temperature is evidenced by the positive responses of a globe thermometer to the radiant field and. B-91 . New York. + Demonstrated operational efficiencies of low-intensity radiant systems established that space comfort control can be maintained with a lower energy input. American Society of Heating. Atlanta. equal comfort conditions are maintained by radiant systems with lowered thermostat settings.nt heating equipment support an overall reduction in the input energy requirement for radiant heating system installations. + The floor heat reservoir reradiates and convects heat into the space. A. ASHRAE Trans. These values are based on 20 years' experience and have been applied successfully in thousand of installations . access to these heating units is restricted for inspections and maintenance. It is in French and appears to be published by technical society in France. This is a design manual for selecting and installing radiant floor heating systems using using plastic pipe. Manufacturers of radiant equipment use adjustment factors of . high-intensity. Correa. (2) porous. Manual Techniques "Avadis". porous. * * Chapot. radiant heaters surpass convective forced-air counterparts in heating large open bays in the following ways: (1) by providing increased thermal comfort at the floor level while substantially reducing heating costs and heat stagnation. S. Edward L. Jean Robert. a malfunctioning burner would be visually apparent by intermittent burner incandescence. Due to the inherent nature of suspending these heaters near the ceiling of an aircraft hangar. (3) when the burners glow a dull red. A.80 to . Design Guidelines for Heating Aircraft Hangars with Radiant Heaters. Pont-A-Mousson. Natural gas may not be available at the hangar location. * * * 5. refractory heaters are safe for use in aircraft hangars. IR radiant heaters are recommended for use in aircraft hangars if natural gas is available because of the following: (1) the porous. Naval Civil Engineering Lab.. The results of this investigation indicate that radiant heaters are practical heaters for use in large open bay buildings.85 in design calculations. B-92 . (2) by being able to heat objects to just above the dew point temperature to prevent condensation and corrosion. 3. It would be useful to have translated if one is going to be interested In installing embedded pipe floor radiant systems. (3) by allowing heating flexibility with zone or whole building heating. Utilization Chauffage Par Le Sol. refractory. Gas fired." * 4. . 2. The gas-fired heaters require ventilation for elimination of the flue gases. IR burners emit heat energy primarily in the longer wavelengths (2 to 6 microns) which is within the optimum absorptance range for personnel and concrete floors and has no adverse physiological effects. December. 1983. refractory. Generally.These factors support the application of an adjustment factor to standard heat loss calculations. The following disadvantages should be noted: 1. 6. 5) Multiple Irradiation. 10) System Design. Verlag C. 8) Design and Construction of Heating Elements. It is not available in English. If this is not done or if the products of combustion are immediately exhausted this percentage reduction does not hold true.. 9) Control Procedures. B. W. 2) Emission of Heat from Large Panels with Imbedded Tubes. 1982. This work is based on the rational heat balance concept and procedure developed in a series of technical papers written by Hutchinson.. 11) Special Radiant Heaters. F. It contains the following chapters: 1) Fundamentals of Heat Transfer. It does have a chapter on panel heating design procedure. Design of Heating and Ventilating Systems. Muller. The Industrial Press. and (4) lower air temperatures for comfort conditions can be maintained with radiant heaters. 1977. F. * * Hutchinson. 1948. The method can be applied to ceiling. 1955. Karlsruhe. Inc. * 7. B-93 . They recommend a minimum roof vent opening of 50 sq.000 Btuh for unvented overhead heaters. area heating and spot heating. NY. Strahlungsheizung Theorie und Praxis. per 100. and 12) Comprehensive Examples.. NY. Radiant Heating .. New York. 7) Heating Requirements and Location of Heated Surfaces. They make several recommendations concerning the application and installation of these types of heaters. . * * Hutchinson. F. 6) Mechanisms of physical Heating in Radiant Heated Rooms. It also allows calculation of design load for the building. Industrial Press.Theory and Practice is a German design book on radiant heating systems. W. The method is simple and is stated to give results with accuracy equal to the analytical procedure. This handbook of fuel gas engineering practices discusses the application of gas fired infrared radiant heaters. * 9. This source indicates that the design heat loss can be reduced by 15 percent if the perimeter heating method is used. New York. This is a textbook for HVAC System design. The method presented here is similar to that given in a book published by Revere Copper and Brass Co. * 8. 4) Heat Emission from Infrared Rays. (2) no air movement required for heat distribution. * * Gluck. Revere Copper & Brass. They also discuss total plant heating. (3) the gas unit efficiency with unvented combustion is about 90% rather than 80% as with vented convection heaters. Germany. 3) Heat Emission from Segments of Surfaces and Radiant Panels. in. A Graphical Design Procedure for Radiant Panel Heating. floor or wall panels. Gas Engineers Handbook. The reasons given for this 15 percent reduction are: (1) very little temperature stratification. building heat loss correction factors. "Infrared Heating for Overall Comfort". etc. A design procedure is presented and some useful information is given for infrared heaters. This is particularly true with factories. Fred J. and acceptance of a lower temperature for comfort. Therefore. Dec. A9. radiant heating. 16 ft Radiant. reduction in the infiltration rate.10. which is based on a combination of the mean radiant temperature and the air temperature. warm ceiling Medium and high temperature radiant units from high levels 0 0-5 0 0 Forced warm air convective system with cross flow at low level 15-30 5-15 0-5 Forced warm air convective system with downward flow from high level 10-20 5-10 0-5 Medium and high temperature cross radiant from intermediate level 5-10 0-5 0 * 11. This combination is based on the resultant temperature being in the range for human comfort. a uniform temperature throughout the height of the heated space is assumed. Useful design charts are presented for heat delivery from heaters at specific locations. ASHRAE Journal. 1968.. In this heat loss calculation. Percentages to be added to the calculated heat loss to allow for these temperature gradients are given below. An environmental temperature is defined. * * Prince. Estimation of Plant Capacity. B-94 . CO2 dilution air requirements. 1975. IHVE Guide. This British guide for designing heating ventilating and air conditioning systems contains some design information for radiant types of heating systems. IHVE. this will also affect the design heat loss calculations. Factors which cause this adjustment are: improved insulation. Method and type of heating " Percent to be added for the following heights of heated space 16 to 32 ft. Certain modes of heating cause vertical temperature gradients which lead to larger losses. particularly through the roof. warm floor Radiant. surface temperature charts. 0 0-5 0 0 >32ft. as adjustments are made in the air temperature. where air and mean radiant temperature may differ appreciably. F. ASHRAE Journal. October. In the interim. A textbook describing the analysis and design of panel heating systems. V. A general description of the selection and application procedures for infrared heaters. 1957. ASHRAE. ASHRAE Research Laboratory. a paper of this length cannot cover all or even a small part of present practices. together with other papers which contain supplementary data. Chapters included are: radiation equations. Room air temperature is the selected criterion of comfort. His conclusions are as follow: "With the growing popularity of gas infrared heat. * * Raber. and Hutchinson. 63. 1947. J. Having assured himself of the validity of such claims. and glass does not differ greatly from room air temperature. Prince. It gives descriptions. * 14. * * Subcommittee of TAC. he should apply sound engineering principles in the design of any infrared system. This paper presents a simplified procedure for the thermal design of water heated floor panels for use in residences and commercial buildings. evaluation of shape factors and heat balance equations. the ceiling.. showed that this near-equality of the 2 temperatures normally prevails.12. Panel Heating and Cooling Analyses. B-95 . The procedures in both papers are based primarily on the experimental data obtained at the ASHRAE Research Laboratory under the guidance of the ASHRAE Technical Advisory Committee on Panel Heating and Cooling. it is suggested the heating engineer evaluate carefully and demand proof of all performance claims regarding infrared units before he makes his selection. and the design procedure is restricted to situations in which the area-weighted average temperature of the walls. It complements the previously published paper. comfort relationship. The room-scale tests. architects and the heating industry need standards for evaluating units and for system design. This work has been reported in a series of research papers which are listed in the references. "Thermal Design of Warm Water Concrete Floor Panels". John Wiley & Sons. F. "Selection and Application of Overhead Gas-Fired Infrared Heating Devices". B. W. mean radiant temperature. With the many thousands of successful installations. Reputable manufacturers will have engineering and application manuals which should be studied. engineers. Trans. 1962.. advantages and disadvantages of panel systems. F. NY.. which simulated various conditions of construction and outdoor temperature. * 13. Since applications have become so varied. previous history on almost any type of application is available and the experience gained should be utilized". A panel designed by this procedure will maintain the desired room air temperature for the selected outdoor conditions. B-96 . Planungsunterlage Fur Ingenieure. The pamphlet is in German and presents design and installation details. . Heizungstechnik Kg.15. Thermo Lutz. Thermolutz GMBH and Co. This is a design procedure presented by a German manufacturer for hydronic floor heating systems. pg. Energy Management Technology. No. B-97 . In the fall of 1976 radiant systems were installed in both buildings.. The purpose of this project was to compare energy consumption for radiant and convective heating units. Correa. An overall comparison for the 2-yr period was made by dividing the total energy consumption by the sum of all the AT's.. The results of measurements of hangar air infiltration and stratification. They were originally both heated convective systems. July 1984. Pt. Discusses the conversion from a forced-air oil-fired heating system to an infrared heating system in an auto dealership. Technical Report R-910. E. vinyl strip doors. The convective systems for both gas and electricity used about 15% more energy/AT than the corresponding radiant system. V. * * . door seals.J. Two buildings on the Rose-Hulman campus were selected -. 7. H. November 1983.one is heated with gas and the other with electricity. ASHRAE Trans. 8. with convective heat. Walker-Davis Publications.. with a radiant heat and 1/2 yr. 86. * 3. "An Experimental Comparison of Energy Requirements for Space Heating with Radiant and Convective Systems". with the electric system radiant heat was used during the first part of the first year and the last part of the second year. One reason for a 2-yr. 1980.. A measure of heating effectiveness is obtained by comparing daily energy consumption with AT where At is the inside temperature minus the average outside temperature for the day. Ashley. They indicate about a 40% savings in fuel costs and a payback of 1 1/2 to 2 years. 44-45. J. For example. Naval Civil Engineering Laboratory. and Canfield. 1. two major causes of heating related energy consumption. Vol. * * Bailey. are reported. vehicle access doors. R. "Energy Conservation: Heating Navy Hangars". and radiant heating) were evaluated and are discussed. "Fuel Bills Halfed After Switch to Infrared Heating System". test was to be able to have each system operate for both halfs of the heating season. Methods to reduce this type of energy consumption (reduction of air infiltration and installation of destratifiers. K. * 2. Design criteria providing hangar air infiltration rates versus hangar size and climatic conditions and design criteria for hangar destratifiers were developed and are presented. How energy is used for hangar heating and what methods are used to reduce hangar thermal energy consumption were investigated. Energy consumption and temperature data were recorded daily for two heating seasons with each building heated about a 1/2 yr. ENERGY CONSUMPTION 1. L. Discussion of a specific building which went through a retrofit and used a gas infra-red heating system to replace a hot water heating system. ASHRAE Journal. The overall efficiency of two heating plants was measured in two well insulated and identical houses. Another building experienced an 10% reduction in fuel usage. The results also show that the values obtained for the electrical heating system are higher than those for the classical system even when this one used low water temperatures. No. and Zito V. pg.New Infrared Heating System Solves Old Problems". Co. Vol. November 1983. Troup Publ. M. Commission of the European Communities. Discusses the installation and operation of a gas radiant heating unit which achieved 90% operating efficiency and saved approximately 50% on energy requirements for the structure. In the first house (heated with a radiant floor) a reduction of 9% of the energy consumption of the boiler was made by using a closed circuit instead of an open one. * * . Edited by Ehringer H.. S60. 48. "Building Heat with Natural Gas Infrared". * * Grum. This authors analysis shows that 15% reduction in unit size as well as energy consumption were present. and a reduction of 11% of the energy consumption was obtained with a night set back from 10 pm to 6 am. . 60-61. "Integration of Different Energy Saving Possibilities in Dwellings". This article presents some heating cost data for particular industrial buildings for various years of operation. They reported a payback of one year and are expecting up to 80% savings in energy costs. and for the second one direct electrical heating with convectors was used. __. September 1978.4. * 6. The Heating and Air Conditioning Journal. "Gas Radiant Heating Installation Achieves 90% Overall Efficiency". Energy Savings in Buildings. Reidel Publ. * 7. B-98 . "In Old Building . It was also claimed that the heating systems provided greater comfort for the workers. June 1968. D.. 1985. * 5. R. The differences of these heating installations are: for the first one radiant floor heating using a high efficiency and low water temperature boiler was used.. * * Guillaume. April. E. Air Conditioning Heating and Refrigeration News. and it was observed that standard-based (65 F) degree-days could be in error by a factor of two. M. V. the warehouse area used off-peak thermal-electric storage units. Nelson. Variable-base degree-days were computed for each zone. humidities. The office area was heated by a heat pump with an auxiliary backup furnace. The warehouse had few internal gains. was monitored for two heating seasons using a computerized data acquisition system. and Pate.. An analysis of the thermal performance of both the building envelope and the heating equipment was performed. A data base has been established that will be used in future studies to verify the loads obtained from computer programs. M. R." "It was observed that each of the three spaces had a different base temperature to be used in the calculation of degree-days and that the use of standard degree-days based on 65 F could lead to erros in heating load estimates. The base temperature in the garage was significantly lower than the thermostat setting because a radiant-type heating system was used. each having different heating equipment. resulting in the degree-day base temperature being close to the thermostat setting. 1. B. "An Experimental Study of a Multipurpose Commercial Building with Three Different Heating Systems". The verified programs will then be used to assess the economic impact of various energy conservation measurements". R. Pt. "A commercial building with three distinct zones. M. Trewin. Langdon.. The zone loads and equipment energy consumption and performance were calculated. the garage area heating load was met by a gas-fired radiant system.. B-99 . R. The microcomputer-based data acquisition system obtained and stored hourly temperatures.. 93. F.8. ASHRAE Trans. 1987. The results for the office space were in good agreement with the65 F based degree-days as expected. and energy flows. Piping & Air Conditioning. * 3. The discussed models of floor heating and oil furnace as well as the overall model of a heated space have proven to be sufficiently accurate and simple. Floor coverings would delay the air change rate. The occupant's response and acceptance of such a plan were investigated. . Rascati. panel. 4" in below the surface laid on 12" centers.W S Messe. "Heating Panel Time Response Study". Markel. The air temperature reached 50% of its change in 4 hours and 90% in 9 hours. The heat input rates were high (5 to 6 times steady state loss) in order to account for thermal storage. energy consumption can be reduced if intermittently occupied spaces are kept at a low. Use a 4" concrete slab laid on 9 in. Ottin. Four hours were required for the panel to reach 63% of its change. The importance of the radiation heat transfer for the accuracy of transient models is recognized. Only a small quantity of results are given and no equations are presented. The floor had 1" OD pipe. * * Berglund. 98. 1982. Fast acting radiant heaters were activated when the subjects entered the cold space bringing the mean radiant temperature to a high level and the operative B-100 . pg. In winter. A floor panel heating system was tested for transient response. TRANSIENT EFFECTS 1. M. March. Fort K. 1949. Aiulfi. "Radiant Assisted Comfort Heating for Energy Conservation in Intermittently Occupied Spaces". 1985. Ireland.. D. Energy conservation in the Built Environment: Proceedings of the CIB W67 Third International Symposium.K.. no windows and a single door. R. * * Algren. * 2... Subjects from a comfortable area at 22oC entered a space at 15oC and occupied it for 2 hours. Ventilating and Air-Conditioning Systems. There was an hour and a quarter lag between the air temperature and the panel surface temperature at the 63 percent value. The MRT was approximately 3oF higher than the air temperature for the majority of the time. The maximum air temperature rise was 8oF per hour and occurred about an hour after the heat input. Finite differences are used for solving the conduction equation. Clima 2000 . A continuous record was made of water.. oil furnaces and rooms are developed for the application of microprocessors for online control strategies. Vol. T. Dublin.Heating. 6. Ben. "Modelization of Floor Heating and Oil Furnaces for the Unilization of Microprocessors in DDC". A. air and surface temperatures resulting from a sudden and large change in the heat input to the panel system. Heating. Room had well insulated walls. Transient Models of floor heating. W S Kongres .. B. L. and Ciscel. ground. energy conserving temperature when unoccupied and raised to a comfortable level when occupied. of crushed rock. For modeling short-term dynamic responses. The technique and control are also applicable to steady state situations. Spot and heated ceiling type radiant systems were tested. degree of comfort and thermal acceptability were gathered periodically during the tests. A sensor that averaged air and mean radiant temperatures was found to be superior to an air temperature sensor as input to the radiant heat controller. The air temperature increased at 3oC/h during the exposure to simulate the response of a conventional convective heating system. The savings depend on the application and can be predicted by calculation from the response characteristics of these tests. Subjective responses of thermal sensation. M. 1982. "Thermal Room Models for Control Analysis". 87. R. particularly where there is intermittent occupancy. Pt. Comparison tests are described that show people will accept spaces being cool upon entry if the spaces can be brought quickly to a comfortable level with radiant heat. * * Borresen. Rascati. Pt. V. V. It is shown that the choice of the simplification level employed depends on how closely the long-term responses and steady-state values are to fit the actual room response. The room air is assumed to be fully mixed. a simple time constant corresponding to the air change rate of the room is usually adequate and will lead to choosing conservative control parameters.temperature or the temperature that the environment feels like to 22oC. L.. * 5. ASHRAE Trans. "Radiant Heating and Control for Comfort during Transient Conditions". L. This paper discusses four simplified dynamic room models which in different ways take into account the thermal interaction between room air and surrounding walls. After the subjects' entry. The 16 subjects judged the environment of the radiantly heated system for intermittent occupancy to be thermally acceptable. B. 1981.. Subjects from a comfortable area at 22oC (72oF) entered a space at 15oF) and occupied it for two hours. * * Berglund. 2. The analysis of a dynamic control loop often requires the use of a room model. The radiant heaters were regulated by operative and air temperature controllers. 2.. The radiant system controlled by operative temperature was more acceptable and more energy efficient than the air temperature controlled radiant system because it produced less overheating. An experimental procedure for determining typical parameter values is discussed. ASHRAE Trans. spot radiant or fast-acting radiant ceiling panels rapidly raised the operative temperature of the space to 22oC. and Markel. A. B-1GL . It produced less operative temperature overshoot and greater occupant thermal acceptability and reduced power consumption. 88... There are numerous applications for energy savings with fast radiant systems. * 4. Eight-hour floor warming appears to merit further study. The radiant heating system consisted of copper tubes embedded in a standard plaster ceiling at 6-in (152-mm) intervals. Appls. 92. "Room Temperature Dynamics of Radiant Ceiling and Air Conditioning Comfort Systems". The practical problems of increasing speed thickness. R. 2. IA-19. * 7. "An Experimental Study of the Transient Response of a Radiant Panel Ceiling and Enclosure". The dynamic response is more than adequate for perimeter areas of buildings. as well as extending the design rpocedures to cover eight hour operation.6. V. March. 1983. * 8. B. * * Pfafflin. It seems that with good insulation an eight hour charge period can produce an acceptable temperature variation throughout the day.. * 9. V. 1977 (NTIS-PB 277 115).. Z. * * Zhang. It is found that the dynamic responses are functions of the means of energy input. J.. cost effectiveness of underfloor insulation and the specification of control systems need attention. 1986. 69. A. 1963. Results of tests conducted under rigidly controlled conditions at the former Electric Space conditioning Institute are shown to support the proposed models. The transient response of a radiant heating system and enclosure was investigated for a range of hot-water supply temperatures and flow rates. .. on Indust. Sept/Oct. Pt. Verification of models previously advanced for description of the active and passive modes of baseboard heating and forced air convection is given. ASHRAE. This article disusses the response time for water carrying radiant panel systems. Boyar. Transient experiments were B-102 . M. and Pate. Vol. "Space Heating Dynamics". The author compared the response time (time to reach 90% of its terminal value) of this system to that of a forced air system. No. This brief note looks at the effect of improved building insulation on the temperature variation of a building heated by off peak underfloor heating. R. "Eight Hour Floor Warming: A Feasibility Study".. and found that they were comparable. which have a relatively high percentage of glass. E. * * Mclntyre. ASHRAE Trans. Trans. England. A carpet increases the downward loss substantially and its use reduces the effectiveness of floor warming. Electricity Council Research Centre. IEEE Trans. Sequences of heating and cooling reponses for the two means of energy input are evaluated by means of incremental forms of the fundamental equations. Capenhurst. D. 5. Increasing the thickness of screed over the heating cable is very beneficial. performed by heating the radiant ceiling and enclosure from a cooled-down condition by using a step change in the hot-water supply temperature. Temperature transients in the water supply and return lines on the ceiling and wall surfaces and in the room air were then monitored for a period of several hours. Results were as follows: the ceiling temperature was uniform. and the room walls were heated by a combination of radiant heat transfer from the ceiling and convection heat transfer from the air. In addition. the transient response of the radiant system was found to be a function of water supply temperatures but not of water flow rate. the air temperature did not lag the wall and floor temperature. the thermal response of the ceiling and enclosure was slow because of the large thermal mass in the ceiling. B-103 . It is hoped that others will further develop the device and that its use will further the research on the effects of radiation on the occupants of controlled space. E. -July. * 2.. The basic principles of the radiometer and its construction are very simple. 82. pp.. 1976. A special radiometer described was developed to measure directional mean radiant temperatures.. ASHRAE Trans. Experience up to this time has shown that the radiometer is very handy and durable and further requires only a few minutes to reach equilibrium. Part I. 67-70.L. D. Jr. and by further investigations it should be possible to develop a complete theory for the thermal sensitivity of the radiometer and hence make its construction still more sensitive. ASHRAE Trans. Vol. B-104 . "Thermal Comfort Measurements". The radiometer also has advantages in the determination of the Effective Radiant Field (ERF) proposed by Gagge. * * Madsen. 1976. Pt. The instrument has been successfully vised in a test room with highly reflective walls. James. V. "The Design Construction and Operation of a Scanning Radiometer for Measurement of Plane Radiant Temperature in Buildings". B. L. "A Radiometer for Environmental Applications". 1949. direct measurement of the predicted mean vote in a given space. A description of an instrument is given along with its design details and its initial performance evaluation. and ability to record the results. ease of use. Maglum. * 3. P. His conclusions were as follows: 1. It appears to have some advantages over other known instruments in convenience. V. ASHRAE Trans. 2. 1. * * Korsgaard. A convection-nulling radiometer which involves the use of a thermoelectric module is described.. The calibration of the radiometer is simple. Piping and Air Conditioning.. Pt. Its accuracy is felt to be "at least as good as other devices in use. The new comfort meter provides a quick. INSTRUMENTS 1. Theodore H. 1969.. Benzinger. 75. T.. W. * 4.. and McNall. * * Braun. and Hill. 82.. L. The author describes a comfort meter which he has developed. Heating. "A New Radiometer Measuring Directional Mean Radiant Temperatures". Vol. Madsen. In order to compare different thermal environments. S. simultaneously and at the same position. 1986. Various industrial applications are discussed. * * Tenney. Oct. Comparison with calculated PMV values based on separate measurements of the thermal parameters in typical environments shows good agreement. Mechanical Engineering. 1. While the comfort meter takes into consideration the heat loss of a person as a whole. and v. B-105 . Ill. In cases where man is exposed to asymmetric radiation. Just as in the case of the comfort meter. The comfort meter measures the thermal effect of ta. * * * 5.. . The author indicates that there are no standard measuring methods for determination of the degree of thermal asymmetry or draft. MRT. the comfort meter gives a better approximation to the PMV value than can be calculated from traditional measurements of the thermal parameters.Industrial Radiation Thermometry". as well as for the reproducibility of thermal environmental measurements in common. 4. * 6. it is still important that the thermal parameters are measured accurately. ASHRAE Trans. A discussion of the types of radiation thermometers which are available and their range or field of applications is given. the instrument has been constructed with the aim of simulating a person's heat exchange with the surroundings.2. V.. 1980. this gives a good reproduction especially under non-steady-state conditions. A useful table is given for choosing the correct thermometer.. "Red Hot and Hotter . Neither are there any measuring instruments on the market which are particularly adapted for these measurements. T. A. to measure the thermal effect produced by changes in the heating and ventilating system. The measuring instrument described here is developed as an aid to fulfilling this need. Inaccuracy of a certain PMV value is due mainly to the fact that in practice it is difficult to state activity level and clothing with great accuracy. "Definition and Measurement of Local Thermal Discomfort Parameters". 86. 5. 3. Pt. L. and hence of the central temperature perception. the discomfort analyzer aims at simulating a person's peripheral temperature perception. R. * 3. E. It was also determined that continuous circulation of the water was important in order to avoid surges in panel temperature. * 2. April. Piping and Air Conditioning. * * Algren. The vertical thermostat location did not appear to affect the results. Proper demand heat rates for the boiler were also illustrated. 1959. Discusses the use of spot cooling.Part II". Snyder. when large masses in the panel were available. B. Large glass areas and solar loads increase the transient response and control problems. * * Hazard. 1953. Piping and Air Conditioning. Best performance was obtained from systems with outdoor thermostats used for circulating water temperature reset. .. The field studies indicate that the water temperature should be reset with the outside temperature. for the workers.. The temperature difference across carpeting was lloF and floor tile was 2oF. Discussions are also presented on instruments to be used for measuring radiant heat. Jr. 1954. Snyder. G. W. . The significant results of this two-year study on floor panels in two factory buildings and a residence indicate the importance of proper design and installation of the basic heating system and the importance of the selection and location of various control elements.M. Locke. On-off and modulating valve action controls provided system stability. B. A. F. A. F. It recommends specific practices in both the operation and type of control system.. These are field studies and it is difficult to extrapolate design data from these results. E. J. R. Heating. radiation shields and ventilation for controlling radiant loads in industrial applications. This paper discusses results of field tests covering various control systems in three different types of construction. ... More of the heating load should be delivered to the perimeter of the system below windows.. Head. It was pointed out that returns should not be combined prior to the three-way mixing valves used for recirculation. The outdoor thermostat used for resetting the water temperature should be located such that it is exposed to the same climate conditions (wind and solar effects) as the structure. S. The main objective was to maintain comfort condition. "Field Studies of Floor Panel Control Systems . CONTROLS 1. B-106 . "Radiant Heat Control in Industrial Plants". February. Jr. Heating. Algren. November. "Field Studies of Floor Panel Control Systems". ASHRAE Journal. . etc. and Hwang. but they are not within the scope of this paper. L. 1973. the comfort standard can be met properly designed to position control. * * * 5. Most commonly used thermostats are sensitive enough to db and MRT to be suitable for use with comfort radiant heat systems. Jr. The electric ceiling cable system has a time constant of 40-50 minutes so that a cycling rate of 2-3 cycles per hour is adequate. This includes all of the systems described here.on" cyclers and large-droop thermostats or temperature transmitters are often warranted. p. a linearization of the above equations within the bounds of small deviations is justified. provided that the cycling rate is adequate and the droop is not too great. 7. When radiant sources are used in factories. However. which responds thermally in a manner similar to people.4. the nonlinear equations are linearized around B-107 .. 81. C. An ideal thermostat. "Simultaneous Control of Temperature and Humidity in a Confined Space . such as "% . a thermostat cycling rate of 6-10 cycles per hour is satisfactory to meet" the proposed ASHRAE unsteady-state comfort criteria. so that no single ideal location can be found. ordinary thermostats are not usually recommended. 4. True proportional control (modulating thermostats) can be advantageous in reducing the effects of cycling and may have advantages in reducing demand charges. Building Science. does not exist. 2. Since temperature-humidity control systems allow usually only small deviations of temperature and humidity from a desired operating point. not withstanding considerably R&D effort by the control industry. most spaces are subject to variations of comfort conditions with position in the spaces and load conditions. ASHRAE Trans. etc. The equations are sufficiently general to take into account external heat loads. T. 39-49. 1973. Even if an ideal thermostat did exist. Fan. "A Manufacturer's View of Radiant Heater Control". Hence. and internal heat and moisture loads within the confined space. .. C . Preston E. loading docks.. N.. Pt. L. waiting platforms. Special techniques. McNall. 1. Vol. etc. . Here comfort is not as important as the alleviation of severe cold. 5. 8.Part I. E. 6. Vol. 3. A pair of nonlinear differential equations which describe the transient behavior of temperature and humditiy in a confined space have been derived from simultaneous material balances of dry air and water along with the enthalpy balance of moist air. For radiant heating sources with time constants greater than 5 minutes. increasing efficiency. Nakanishi. Pereira. The author's conclusions are as follows: 1. * * Walker.a desired steady state operating point. 1962. "Control of High Intensity Infrared Heating". and the thermostat control systems. This article presents a short description of various means of controlling infrared heating systems. November. * 6. ASHRAE Journal. He discusses the use of automatic controls vs manual and the use of electronic.. A. The response from the linear equations is found to compare very favorably with that from the original nonlinear equations. On using available relations for the specific volume and enthalpy of moist air. the linearized equations further result in a pair of linear uncoupled differential equations. C. B-108 . N. SPOT HEATING & COOLING 1. Best, W. H., "Spot Heating and Condensation Control Using Gas Infrared Systems", ASHRAE Journal, June, 1968. This article describes the design procedure that might be used for spot radiant heters for comfort control and condensation control. It uses gas infra-red heaters for the specific units. It contains a useful design figure for the energy per hour per ft^ to be applied and a few other general rules of thumb. * 2. * * Olesen, B. W. and Nielsen, R. , "Radiant Spot Cooling of Hot Working Places", ASHRAE Trans.. V. 87, Pt. 1, 1981. Radiant spot cooling can improve the thermal conditions in warm working environments. Radiant spot cooling decreases discomfort caused by warmth, but may create discomfort caused by radiant asymmetry. It is important to optimize panel positions according to the angle factor between worker and panels. Water condensation on the cooling panels and positioning of the panels may in practice limit the use of radiant spot cooling. The efficiency of radiant spot cooling is rather poor (10-15%). A good section on definitions is presented. * 3. * * Sofrata, H. M., and Al-Hukail, Y., "Spot Cooling System Design", ASHRAE Jnl.. Jan., 1987. An interactive computer program for spot cooling system design is discussed. Input includes ambient conditions for the industrial environment, the metabollic heat production and clothing value of workers in the target area, the jet and target area geometry, and the maximum and minimum of the conditioned air at the target area. Their results have been compared with available data in the literature and a good agreement has been achieved. B-109 APPENDIX C LISTING OF COMPUTER PROGRAM C-l LIST OF INPUT VARIABLES CP HF COOL - Presence (True) or absence (False) of ceiling panels - Presence (True) or absence (False) of a heated floor - Presence (True) or absence (False) of cooling panels HIUP HIDOWN EPSI FP Outside ambient temperature Length, breadth and height of the room U-factors for each room surface Standard convection coefficients (incl. radiation effect) for each room surface ASHRAE standard coefficient for upward and downward convection Emissivity of each room surface Person-to-room surface shape factors P CFM ACH HREF SLOPE Q2L Number of persons Supply air cfm/sq. ft of floor area Infiltration air changes per hour Reference height and gradient for the air temperature gradient Sensible heating load due to lights RATIO AMI EFF CLO FCL V RH Ratio of radiative to Dubois area for a person Metabolic rate and mechanical efficiency Clothing level and clothing factor Relative air velocity Relative humidity XPREF NCP NPAL XCP,YCPALCP BCP EPSIP XMULT - Required temperature of panels Number of panels Number of panels in a lengthwise row Coordinates of the center of the panel Length and width of the panel Emissivity of panel surface Multiplier for convection off the panels TOUT ALTH BTH HT U HI XIN PTOL MAXIT - Initial temperature for calculation - Tolerance and - maximum number of iterations for panel temperature convergence Note: A sample input data file is shown at the end of the computer program listing. C-2 C******** FORTRAN CODE irk************* C TRIAL'3 , DEC 4 C C C C IMPLICIT REAL*8(A-H,0-Z) DIMENSION WK(5000),X(10),PAR(10),F(10),X0(10),VAR(25) DIMENSION HIIN(6),OUT(25,25), OUTl(25,25) DIMENSION XCPIN(25),YCPIN(25),ALCPIN(25),BCPIN(25) C CAN GO UPTO 15 PANELS WITH THIS EXTERNAL FCN C C LOGICAL CP,COOL,HF C CHARACTER*21 TEMP(6) CHARACTER*50 TITLE CHARACTER*45 POUT(50), POUT1(50) C COMMON /OUT/ TOUT COMMON /CEL/ XCEL(ll) COMMON /COMF/ AM1,EFF,AICL,FCL,HC,V,PA,RH COMMON /AIR1/ P,CFM,AICFM,ACH,UQ2P,Q2P,UQ2L,Q2L COMMON /Q/ Q1,Q2,QSTD3,QACT3,Q3,Q4,QP5,Q5,Q6,Q7,Q8 COMMON /QNET/ QNET1,QNETP2,QNET2,QNET3,OUA COMMON /QP/ QCVP,QRP COMMON /QI/ QR(6),QCV(6),QCD(6) COMMON /U/ U(6) ,HI(6),CI(6),EPSI(6) COMMON /UP/ UP,HIP,CIP,EPSIP,XMULT COMMON /CONV/ HIUP,HIDOWN COMMON /FSURF/ FS(6,6) COMMON /FPEOP/ FP(6) COMMON /GRAD/ HREF,SLOPE COMMON /TERM/ TERM6,TERM7,ALHS COMMON /DIMEN/ ALTH,BTH,HT COMMON /AREAS/ RAREA(6) C /AREAS/ APPEARS IN MAIN,FCN AND SHAPE PROGRAMS. COMMON /PAN1/ CP,COOL,HF COMMON /PAN2/ PRAREA,PAREA COMMON /PAN3/ NCP COMMON /PAN4/ FSCP(25,25) COMMON /DIMP/ XCP(25),YCP(25),ALCP(25),BCP(25) C C DATA PAR/10*0.DO/ C C C TEMP(1)='TEMP OF FLOOR ' TEMP(2)='TEMP OF CEILING * TEMP(3)='CLOTHING SURFACE TEMP1 TEMP(4)='ROOM AIR TEMP* TEMP(5)='MEAN RADIANT TEMP* TEMP(6)='SUPPLY AIR TEMP' C TEMP(6) IS SUBSTITUTED BY PANEL TEMPERATURE FOR (CP) C C C READ (5,*) XIN READ (5,*) TOUT C-3 c c READ READ READ READ READ READ READ READ READ READ READ READ READ READ READ (5,*) (5,*) (5,*) (5,*) (5,*) (5,*) (5,*) (5,*) (5,*) (5,*) (5,*) (5,*) (5,*) (5,*) (5,*) RATIO AM1.EFF AICL.FCL V RH P CFM,ACH HREF, SLOPE Q2L (U(K3),K3=1,6) (HIIN(K3),K3=1,6) HIUP.HIDOWN (EPSI(K3),K3=1,6) ALTH,BTH,HT (FP(K3),K3=1,6) READ READ READ READ (5,*) (5,*) (5,*) (5,*) N.NSIG, ITMAX TITLE NVAR (VAR(I),I=1,NVAR) C CP IS TRUE MEANS THAT CEILING PANELS ARE BEING USED C IF CP IS FALSE, THE PROGRAM ASSUMES CONVECTIVE HEATING/ COOLING READ (5,*) CP READ (5,*) HF IF (CP) THEN READ(5,*) COOL C COOL IS TRUE MEANS THAT COOLING AND NOT HEATING IS BEING PERFORMED. C THESE 1002 C READ(5,*) XPREF READ(5,*) XMULT READ(5,*) PTOL ,MAXITP READ (5,*) HREF, SLOPE WILL OVERRIDE THE VALUES READ-IN PREVIOUSLY . READ(5,*) NCP READ(5,*) NPAL READ(5,*) EPSIP DO 1002 J = 11,NCP+10 READ(5,*)XCPIN(J),YCPIN(J),ALCPIN(J),BCPIN(J) CONTINUE ENDIF ASDF C C C CCCCCCCCCCCC CCCCCCCCCCCCC 3525 C C C C C C C C C C DO 1550 IVAR = 1,NVAR ACH = VAR(IVAR) READ (5,*) ALTH,BTH,HT READ (5,*) (FP(K3),K3=1,6) DO 3525 J = 11,NCP+10 READ(5,*)XCPIN(J) ,YCPIN(J) ,ALCPIN(J) ,BCPIN(J) CONTINUE READ(5,*) XPREF READ (5,*) (U(K3),K3=1,6) READ(5,*) NCP READ(5,*) NPAL C-4 ! C C C WRITING THE INPUT DATA... C WRITE (6,123) C WRITE (6,123) C WRITE (6,123) C WRITE (6,123) C WRITE (6,234) 123 FORMAT( //) 234 FORMAT(1X,T15,100('*')) C WRITE (6,77) 77 FORMAT (1X/,T5,'INPUT DATA',///) C WRITE(6,78)ALTH,BTH,HT 78 F0RMAT(1X/,T5,'R00M DIMENSIONS:',T25,'LENGTH = '.F6.2, & T45,'BREADTH = *,F6.2,T65,'HEIGHT = *,F6.2) C WRITE (6,66)TOUT - 460.DO 66 FORMAT (1X/,T5,'OUTSIDE TEMPERATURE = ',F6.2,' DEG. F 1 ) C WRITE (6,1)AM1,EFF 1 FORMAT (IX,/,T5,'METABOLIC RATE PER UNIT DUBOIS AREA = ', &F5.1,' KCAL/HR.(SQ.M)*,10X,'EFFICIENCY= ',F4.2) C WRITE (6,2)AICL,FCL 2 FORMAT (IX,/,T5,'AICL = ',F4.1,' CLO',12X,'FCL = ',F5.2) C WRITE (6,3)V 3 FORMAT (1X,/,T5,'V = ',F6.2,' M/S') C WRITE (6,41)RH 41 FORMAT (1X,/,T5,'R.H. = \F4.2) C C WRITE (6,5)P 5 FORMAT (IX,/,T5,'NUMBER OF PERSONS =',F4.1) C WRITE(6,921) RATIO 921 FORMAT(1X/,T5,'RATIO, OF RADIATION AREA TO DUBOIS AREA = ',E15.8) C WRITE (6,8)UQ2L 8 FORMAT (IX,/,T5,'SENSIBLE HEAT (LIGHTS) BTU/HR =',E15.8) C WRITE (6,6)CFM,ACH 6 FORMAT (IX,/,T5,'SUPPLY AIR CFM PER SQ.FT = ',E15.8,10X, &'TOTAL INFILTRATION AIR CHANGES PER HOUR = *,E15.8) C WRITE (6,1003)HREF,SL0PE 1003 FORMAT(IX,/,T5,'USING A GRADIENT FOR THE TEMPERATURE OF AIR', &' AT DIFFERENT HEIGHTS :*,/,T5,'REFERENCE HEIGHT IN FT = ',F6.2, & 5X,'SLOPE (DEG. F PER FT) = ',F6.2 ) C WRITE(6,211)(K3,K3=1,6) C WRITE(6,21) (U(K3),K3=1,6) C WRITE(6,22) (HIIN(K3),K3=1,6) C WRITE(6,235) (EPSI(K3),K3=1,6) C WRITE(6,609) (FP(K3),K3=1,6) 211 FORMAT(IX,//,T5,'SURFACE (I) ',2X,6(I2,14X) ) 21 F0RMAT(1X,//,T5,'U(I) ',2X,6(E15.8,2X) ) 22 F0RMAT(1X,//,T5,'HI(I) ',2X,6(E15.8,2X) ) 23 F0RMAT(1X,//,T5,'CI(I). ' ,2X,6(E15.8,2X) ) 235 F0RMAT(1X,//,T5,'EPSI(I)',2X,6(E15.8,2X) ) 609 F0RMAT(1X,//,T5,'FP(I) *,2X,6(E15.8,2X) ) C C C C C C WRITE (6,31)N,NSIG,ITMAX C C C WRITE(6,123) C WRITE(6,123) C-5 123) C CALCULATING CI FROM GIVEN U AND STANDARD HI C DO 51 Kl=l. CP ? = *. C CIP = l/( 1/UP -1/HIP ) C ENDIF C WRITE(6.6 CI(K1)= l/( 1/U(K1) -1/HIIN(K1) ) 51 CONTINUE C C C C INITIALISING THE UNKNOWNS .1022) 1022 F0RMAT(1H1.NOT.0 X(l)= XIN C TEMPERATURE DISTRIBUTION INITIALLY GIVEN TO THE IMSL DO 20 11=2.XPREF-460.*)' HEATED FLOOR WRITE(6. HF ? = '.DO C-6 .10 X(I1)=X(I1-1)+1.D0 20 CONTINUE C C C C C FOR THE PANELS C WRITE(6.781) XPREF.123) WRITE (6..*)' COOLING WRITE(6.*) TITLE WRITE(6.(CP) IF(HF) THEN TEMP(1)='CEILING TEMP ' TEMP(2)='REST OF FLOOR TEMP ' TEMP(6)='HEATED FLOOR TEMP ' ENDIF C ITERP = 1 C IF (CP) THEN X(10) = XPREF ENDIF C C THIS IS NOT USED RIGHT NOW. COOL SUBROUTINE IS: C TEMP(6)='CEILING PANEL TEMP' TEMP(2)='REST OF CEILING TEMP ' C TEMP(6) IS SUBSTITUTED BY SUPPLY AIR TEMP.10 11 X(J1)=0.123) ? = './) WRITE(6.*)* CEILING PANELS WRITE(6.123) WRITE (6. DO 11 Jl=l. C IF(CP) THEN.123) WRITE (6. PANEL CONDUCTION BEING EXCLUDED FROM THE C ANALYSIS . FOR .C WRITE(6. 2.F7.HT.F6.F6.YCP(J).13 ) WRITE(6.T90.T50.2.MAXITP 782 FORMAT(1X/.NCP.'PANEL CONFIGURATION :'.T20. C & T5.FSCP) C C RESULTS ARE PRINTED AFTER THE XREF TEMP IS CONVERGED TO. CONVECTION COEFF.174) NCP 174 FORMAT(1X/.FS.BCP(J) 176 FORMAT(1X/.T5. & T50.T45. &' BEG.BTH.3.XMULT = '.F6.E15.E15.L C ELSE C-7 .2. F7.D0 DO 1019 J = 11.F6.176)J. &' ITERATION :'.3.'OVERALL U FACTOR (UP) = *.'LENGTH'.8./) C C WRITE(6.'TOLERANCE = '. R OR '.'NUMBER OF PANELS IN A LENGTHWISE ROW = '.FS) C C C THE GOTO STATEMENT LEADS TO THE FOLLOWING 777 777 CONTINUE TALCP = 0.8.//.13 ) WRITE(6.'EMISSIVITY OF PANELS (EPSIP) = '. C & T5.T5.T15.YCP.//. Y (ALONG BREADTH) '.T75.2.T5.T5.T30 'X (ALONG LENGTH) '.ALCP(J).BCP.'BREADTH') C C C INITIALISING THE PANEL GEOMETRY AND DIMENSIONS C DO 3501 IPAN = 11.T45.3.1511)XMULT 1511 FORMAT (//.811)EPSIP C605 FORMAT(1X/.8.'PANEL TEMPERATURE DESIRED = '.T90.15 ) G WRITE(6.175) 175 FORMAT(1X/.T5.T75. NCP+10 XCP(IPAN) = XCPIN(IPAN) YCP(IPAN) = YCPIN(IPAN) ALCP(IPAN) = ALCPIN(IPAN) BCP(IPAN) = BCPIN(IPAN) 3501 CONTINUE C C C C DO 165 J =11.T30. &/.BTH.1023) NPAL 1023 FORMAT(1X/.'MAXIMUM NUMBER OF ITERATIONS = '.F6.HT.'PARAMETERS FOR PANEL TEMPERATURE'.E15.'PANEL DIMENSIONS (INITIAL)f.T5./) 811 FORMAT (//.'COORDINATES OF CENTER'.//) 165 CONTINUE C CALL SHAPE(ALTH.I3.781 FORMAT( 1X/.'TOTAL NUMBER OF CEILING PANELS = '.F6. 'STD.'PANEL NUMBER' &T30.* DEC F*) WRITE(6.(HIP) = '. 'EMISSIVITY (EPSIP) = '.ALCP.NCP+10 WRITE(6.XCP. &//./) WRITE(6.'MULTIPLIER FOR CONVECTION .//.T75. NPAL + 10 1019 TALCP = TALCP + ALCP(J) UALTH = ALTH/ NPAL C CALL SHCP(ALTH.T5.2.T5.2./.XCP(J).782)PTOL.F6. T15.HT.123) C WRITE(6..K) .2X) ) 404 CONTINUE C ENDIF ..BTH.10 Cll X(J1)=0.X.405)1.N.6 WRITE (6.123) C WRITE(6. C & (X(JK).N.FS) C SHOULD LATER TRY TO PASS FS THROU1 THE SUBROUTINE.6(I3.ITMAX.K=1.T5.0 C X(l)= XIN C ABOVE IS NEEDED BECAUSE THIS LOOP IS EXECUTED MANY TIMES.PAR.6) 405 FORMAT (1X/.T15. & 'WITHOUT CEILING PANELS '.*) 'TEMP.'WALL-TO-WALL SHAPE FACTORS'.LINEAR EQUATIONS*********^^^ C CALL ZSPOW (FCN.IER) C TO WRITE OUT THE F AT CONVERGENCE OF EACH EQUATION CALL FCN(X.FNORM.I CALL SHAPE(ALTH.403) (K .RATHER THAN C THROU* THE COMMON WRITE (6.10) C C WRITE(6.6) 403 FORMAT (1H1.I3.K=1.D0 C20 CONTINUE C IF (CP) THEN C X(10) = XPREF C ENDIF C C C C C C WRITE(6.WK.JK=1.l) WRITE(6./.9X) ) C DO 404 1=1.//.T35.(FS(I.PAR) C IER IS WRITTEN LATER ON C C WRITE(6.123) C-8 .EQ.10 C X(I1)=X(I1-1)+1. DISN.F.4.__1_^^__1_^1_I_1__J_1_1^^ C c c c c c c c 606 CONTINUE -• C INITIALISING THE UNKNOWNS .AFTER FIRST ITERATION'.6(E10.123) C C***A*****A*S0LOTI0N OF THE NON.5X. C DO 11 Jl=l.123) C IF(ITERP.NSIG. C TEMPERATURE DISTRIBUTION INITIALLY GIVEN TO THE IMSL SUBROUTINE IS: C DO 20 11=2. ITERATION IS NOT'.EQ. C &' ITERATIONS' ENDIF C ENDIF C C C =========================^^ c C C WRITE(6./) WRITE (6. C WRITE(6.123) WRITE(6.IER WRITE(6.*)* X(10) AT '.T90.*)'IER = 130 : CANNOT GET ACCURACY (NSIG) REQUIRED * WRITE(6.176)J.'COORDINATES OF CENTER'.T75.' = \ALCP(J) ELSE C THIS IS FOR PANEL HEATING. PTOL . MAXITP) THEN C WRITE(6.BCP(J) 786 CONTINUE C-9 .EQ. &/.*).YCP(J).T35.123) IF(IER.'BREADTH') DO 786 J =11.234) WRITE(6.GT.T15.'PANEL CONFIGURATION :'.J.'PANEL NUMBER'.*)'IER = 129 : NUMBER OF CALLS TO FCN HAS EXCEEDED*.ALCP(J) ENDIF 778 CONTINUE ITERP = ITERP + 1 GO TO 777 ELSE C IF(ITERP.*)'ALCP(.ALCP('.*) ' ERROR PARAMETER OF IMSL ROUTINE..OR.' = *. & ' ITMAX*(N+1).X(10) DO 778 J = ll.123) WRITE (6.1021) 1021 FORMAT(1X/.NCP +10 IF(COOL) THEN C THIS IS FOR PANEL COOLING .123) OF PANEL CASE. C***** PRINTING OUT RESULTS WRITE(6.ITERP. & T50.'LENGTH'.J.fY (ALONG BREADTH) '.234) WRITE (6.ITERP.ITERP.T75.' = '.GT.*)' NO CONVERGENCE IN'.' ) AT '. ITERP.MAXITP.//.88) 88 FORMAT (1H1. IER.T5.IER = '.' ITERATIONS' ELSE C WRITE(6.ALCP(J).INCREASE AREA IF COOLER THAN REFERENCE ALCP(J) = ALCP(J)* XPREF/X(10) C WRITE(6.*)*IER = 131 : MAY TRY NEW GUESS./ / C C C WRITE(6.MAY TRY NEW GUESS* WRITE(6.ITERP.NCP+10 WRITE(6.130 . 131) THEN WRITE(6. INCREASE AREA IF HOTTER THAN REFERENCE ALCP(J) = ALCP(J)* X(10)/XPREF .*)' CONVERGENCE IS OBTAINED AFTER '. &T30.T30 'X (ALONG LENGTH) *.'OUTPUT DATA '.' ) AT '.LE.AND. & ' MAKING GOOD PROGRESS' c #mmmmm# mmmmmmm AN ABRUPT STOP GO TO 999 ENDIF C IF (CP) THEN IF ( DABS(X(10)-XPREF) .XCP(J).'PANEL DIMENSIONS (FINAL) '.MAXITP) THEN WRITE(6. (T5.T15.6 WRITE (6.' COMPUTED LENGTH OF PANELS EXCEEDS '.*FSCP('. & 'WITHOUT CEILING PANELS '.J). &'THE LENGTH OF THE ROOM*.T5.192) (K .E10.I3./) DO 201 K = 1./.F10.//.K=1.J. USED FOR PANELS = '.K=1.8) 812 C C C C PRAREA IS NOW AVAILABLE THROUGH THE CALL SUB.T15.' .1511)XMULT WRITE(6.(T5.\ WRITE (6.9X) ) DO 193 1=1.9X) ) C 213 212 C C ='.T15.6(E10.6(I3.F10.6) FORMAT (1H1.6) FORMAT (1X/.811)EPSIP WRITE(6.(FS(I.FSCP(K.T15./.4.195) FORMAT (1X//.207) PRAREA 207 FORMAT(1X/. TALCP.785) PAREA F0RMAT(1X///.T5.6 WRITE (6.T5.1018) 1018 F0RMAT(1X///.K) . INITIALLY') WRITE(6.GE.T5. SQFT 192 194 193 195 199 198 413 412 201 C 411 WRITE (6. UALTH .I3.6(E10.OR.K).199) (J.') = '.FSCP(J. &' PANEL WIDTH .(FSCP(I./) ) CONTINUE WRITE(6.412) (K.T5.13.413) FORMAT (1X//.'PANEL-TO-WALL SHAPE FACTORS (FSCP) '.234) WRITE(6.ALTH) THEN WRITE(6.194)I.K) .5X.') = \E10.4..'FSCP(P.'WALL-TO-PANEL SHAPE FACTORS (FSCP) '.(BCP). J=11.6) FORMAT (1X/.K=1.T5.P'.T5. WRITE(6.T5.K=1.PRAREA.'CHOOSE A GREATER VALUE OF'.I3.6) FORMAT (1X/.E15.'TOTAL PANEL AREA =*.'CONVECTION COEFF./) ) CONTINUE WRITE (6.'WALL-TO-WALL SHAPE FACTORS (FSCP) '.2X) ) CONTINUE ENDIF C THIS IS THE END OF THE IF(CP) C C CALCULATING OPERATIVE TEMPERATURE AND ERF C C-10 .411) (K .6) F0RMAT(1X/.213)1.NCP+10 WRITE(6.'WALL-TO-WALL SHAPE FACTORS *.K.2) DO 212 1=1.234) C$$$$$$$$$$$ AN ABRUPT STOP $$$$$$$$$$$$$$$$$$$ C GO TO 999 785 C?$$$$$$$$$$$$$$$$$$$$$$$$$$$$$$$$$$$$$$$$$$$$$$$$ ENDIF WRITE(6.4.I3.6(I3.I3. &//./) DO 198 J = 11.6 WRITE(6.GE.4.812)HIP FORMAT (T5. K=l.NCP+10 ) FORMAT(1X/.T35.T35.'REMAINING CEILING AREA.2) IF (ALCP(ll) .2X) ) CONTINUE WRITE(6. 3X.123) C WRITE(6.I2.PRAREA TOTAT = TOTAT + PRAREA*X(6) ELSE TRAREA = TRAREA + RAREA(6) TOTAT = TOTAT + RAREA(6)*X(6) ENDIF C THIS WILL GIVE AUST IN RANKINE .'.5D0) TOCEL = (HRSI*XCEL(9) + HCSI*XCEL(8) )/(HRSI+HCSI) TOF = TOCEL*1.234) C WRITE(6.D0 HCSI = 8. P8.12.XCEL(J).8.T5.5D0* (V**0.67D-8 XCOMF = ( XCEL(8) + XCEL(9) )/2.'='.'X('.234) C C C C WRITE(6.8.123) WRITE(6.123) C WRITE(6.5 TOTAT = TOTAT + RAREA(I)* X(I) TRAREA= TRAREA + RAREA(I) 1004 CONTINUE C IF(CP) THEN TRAREA = TRAREA +.D0 HRSI = 4. AUST TOTAT = O. C ) * . & E15.15 W/SQ.2. SIGSI = 5.'FNORM = \E15.123) C WRITE (6.D0.SQ.1X.123) C C WRITING THE INITIAL TEMPERATURE CHOICE WRITE(6.101)XIN -460.100)J.DO TRAREA = O.4 (THE FOLLOWING ARE IN SI UNITS.T40.J. 5 TO 10 C-ll .8D0 + 32.123) C THE TEMPERATURES ON OUTPUT C TEMPS. *) ='.2.15D0 C C C CALCULATING AVERAGE UNHEATED SURFACE TEMPERATURE. REFER ASHRAE(F)1985 .5X.'X(1) INITIALLY ='.F6.J..D0)/1.T5.FT ERFFPS = ERFSI/3.E15. CELSIUS.8D0 C C C C WRITE (6. & T88.DO DO 1004 I = 1.8) WRITE(6.T75. 'F(' .T15.///.'TEMP OF WALL1.*)' (NOTE: TEMPERATURES ARE CALCULATED IN THE PROGRAM IN &DEG.4 WRITE(6.(.FN0RM 101 F0RMAT(1H1. RANKINE ) ' WRITE(6.M = 1 BTU/HR.C C C RATIO IS READ IN.123) C WRITE(6.D0. X(J)-460.F(J) 100 FORMAT (IX. AUSTC IN DEG.D0 + XC0MF)**3.DO* SIGSI*RATIO*(273.12./) 10 CONTINUE C AND TEMPS. ')=*. 1 TO 4 DO 10 J=l.XCEL(8) ) 3.DO ERFSI = HRSI*( XGEL(9) . AUST = TOTAT / TRAREA AUSTC = (AUST-492.F6. T5.J5.T5.906) Q5 + Q3 906 F0RMAT(1X//.8./) WRITE(6. X(J5)-460. & '= '.1X.£15.'T0F= '.EIS. & ' AND Q-DESIGN-OVERALL = '.E15.' DEGF.' %'.922) TOCEL.F(J5) 206 FORMAT (1X.8.1514) PDIFF2 1514 FORMAT(1X//./) WRITE(6.T5. & ' AND Q-DESIGN-C0ND2 = \/. & * PDIFF1 = -((Q7 + QSTD3) . & '= '.T5.D0/(Q7+QSTD3) WRITE(6. & ' DEG.2.Q-DESIGN = Q7 +QSTD3 = '. X('.' BTU/HR. & * W/SQ.D0/(Q7+QSTD3) WRITE(6.8) PDIFF1 = -((Q7 + QSTD3) .(Q6 + QACT3) )*100.8. AUST : & * = \E15.F6.12.8) C CALCULATING % LOAD OF DESIGN PDIFF3 = -((Q7 + QSTD3) .A21.M *.12.T5./.'AVERAGE UNHEATED SURFACE TEMPERATURE. & ' PDIFF2 = -((Q7 + QSTD3) ./.T5.(Q5 + Q3) )*100.' DEG.' DEG.1513) Q5 + QACT3 1513 F0RMAT(1X//.E15.D0/(Q7+QSTD3)'.*ERFFPS = '.8.*PERCENTAGE DIFFERENCE BETWEEN STD.8.*('.902) QNET1 902 FORMAT(1H1.'PERCENTAGE DIFFERENCE BETWEEN STD.T5.5X.T5.T5.2.(Q5 + QACT3) )*100.2.*Q-DESIGN-0VERALL = Q6 +QACT3 = \E15.'Q-DESIGN-C0ND = Q5 + QACT3 = *.907) Q7 + QSTD3 907 F0RMAT(1X//.' %'.8) C CALCULATING % LOAD OF DESIGN PDIFF2 = -((Q7 + QSTD3) .F7.T5.'QNET1 = Ql +Q2 -Q3 +Q4 = '.T40.T5.903) QNETP2 903 FORMAT(1X//.905) Q6 +QACT3 905 F0RMAT(1X//.*PERCENTAGE DIFFERENCE BETWEEN STD.') ='. & ' AND Q-DESIGN-COND1 = '. & ' PDIFF3 =.15X.'EFFECTIVE RADIANT FIELD . -((Q7 + QSTD3) .F7.'QINPUT = Ql +Q2 = '.206)J5.8) WRITE(6.(Q1+Q2 ) )*100. & '= '.Q-DESIGN*. C'. C )'.E15.923) ERFSI3ERFFPS 922 FORMAT(1X/.XCEL(J5). F') 923 FORMATQX/.F7.904) QNET3 904 FORMAT(1X//.8) WRITE(6.'STD.123) WRITE(6.2.'F(*.D0 1005 FORMAT(1X//.Q-DESIGN'.E15.T5.AUST-460.DO 13 J5=5.SQFT./) 13 CONTINUE C WRITE(6.'OPERATIVE TEMPERATURE : TOCEL = '.(Q5 + Q3) )*100.'.'Q-DESIGN-C0ND = Q5 + Q3 = '. & T88.*) C WRITE(6.Q-DESIGN'.1031) PDIFF3 1031 F0RMAT(1X//.'QNET3 = Ql +Q2 -Q3 -Q6 = '.15X.8) WRITE(6.D0/(Q7+QSTD3)'./.3X.TEMP(J5-4).' %'.E15. ') ='.1006) PDIFF1 1006 FORMAT(1X//.1005) AUSTC.'QNETP2 = Ql +Q2 -Q3 -QP5 = \E15.D0/(Q7+QSTD3) WRITE(6.E15./) WRITE(6.T5.ERFSI = '.TOF WRITE(6.T15.ISX.E15.T5.(Q5 +QACT3) )*100.T5.T5.') C C C C C C WRITE(6.D0.10 WRITE(6.8) WRITE(6.T5.'='.908) Ql + Q2 908 FORMAT(1X//. C.E15.S. & E15.D0/(Q7+QSTD3) C-12 .D0/(Q7+QSTD3)'.E15.8) PDIFF4 = -((Q7 + QSTD3) .8.E1S.T5.(Q6 + QACT3))*10O.8. 2.'QPANEL/(RAREA(6)*(X(10)-X(8) )) = *.T5.8) WRITE (6.' DEG.8.'PARM4 = '.'='. X(10)-460./.1034)Q1/RAREA(6) FORMAT(1X//.SQFT.T5./.QRP 806 FORMAT(1X//. &'Q3= ACTUAL INFILTRATION LOSS = 1.1037) PARM3 FORMAT(1X//.'Q2= HEAT INPUT BY PEOPLE AND LIGHTS'.202) Q1.D0/(RAREA(6)) 1033 FORMAT(1X//./.806)QCVP.Q3 FORMAT (IX .1515) QACT3 1515 FORMAT (1X.D0 F0RMAT(1X//.'='.T5.//. & 809 & & 1034 & 1035 & 1036 & 1037 & & 1038 & & ' = ' T**7 0 ' *Y ' ^ WRITE(6.QCVP = '. 202 C-13 .8) WRITE(6.'='.809)Q1/PAREA .'HEAT OUTPUT BY PANELS DOWNWARDS. &'Q1= NET HEAT INPUT TO THE ROOM BY SUPPLY AIR OR PANELS'. ' FOR PANELS AT '. ' QPANEL/( PAREA*(X(10)-X(8) )*(X(10)**4-AUST**4)*SIGMAR ) ' . &'QSTD3 = STANDARD INFILTRATION LOSS = 1.911) QSTD3 911 FORMAT (1X.T65.8. & '= '.1033) PAREA*100.'PARM3 = '.8.F10.2. T65.') PARM1 = Q1/(PAREA*(X(10)-X(8) )) WRITE(6./. T55.Q1/RAREA(6) = '.E15. ' AREA OF WALL.6.QRP = '.T5./.1512)100. PER UNIT'. & T65.E15.F10.' BTU/HR.'='.F7.'PERCENTAGE OF CEILING COVERED WITH PANELS*.E15.T5. &'QACT3 = INFILTRATION LOSS (NO GRAD)= 1.(Ql + Q2) )*100.T5. & F7.1032 WRITE(6.SQFT.08*CFM*(TA-TOUT)'.//. & T75.E15.1714D-8 PARM3 = Ql/( PAREA*(X(10)-X(8) )*(XC10)**4-AUST**4)*SIGMAR) WRITE(6.1036) PARM2 FORMAT(1X//.T5. &T5.'CONVECTION FROM PANELS TO ROOM .2.T5.'PERCENTAGE DIFFERENCE BETWEEN STD.08*CFM*(75-TOUT)'.'QPANEL/( PAREA*(X(10)-X(8) )) = '.E15.T5. ' AREA OF PANELS.T5.1038) PARM4 FORMAT(1X//.08*CFM*(XINF-TOUT)'.E15. & E15.'HEAT OUTPUT BY PANELS DOWNWARDS.Q-BESIGN *./) IF(CP) THEN WRITE(6.Ql/PAREA = '.T5. T65.Q2.' %'.T5.1035) PARM1 FORMAT(lX//.8) PARM4 = Q1/(RAREA(6)*(X(10)-X(8) )*(X(10)**4-AUST**4)*SIGMAR) WRITE(6. T55.E15.'PARM2 = f.1032) PDIFF4 FORMAT(1X//.'PARM1 = *.' % ' ) WRITE(6.8) C C C-— ENDIF C WRITE (6.T5.' BTU/HR.T5.2.DO*QRP/(Q1+Q2) 1512 F0RMAT(1X//.8) SIGMAR = 0.F ') WRITE(6.'PERCENTAGE RADIATION = 100* QRP/(Q1+Q2) = '. & ' AND Q-INPUT= Q1+ Q2 = '.D0/(Q7+QSTD3)'.F7.T5.8) PARM2 = Q1/(RAREA(6)*(X(10)-X(8) )) WRITE(6.' .'RADIATION FROM PANELS TO ROOM. ' QPANEL/(RAREA(6)*(X(10)-X(8) )*(X(10)**4-AUST**4)*SIGMAR ) ' .T5.E15.2. PER UNIT'. & ' PDIFF4 = -((Q7 + QSTD3) . &T75.E15.//.8) WRITE (6.T5. T20.801) Q4 FORMAT(1X//.E15.234) WRITE(6.'A POSITIVE VALUE DENOTES A LOSS FROM THE SURFACE'.'='.T8.T5.E15.T35.'='.E19. & T65.fQR(I)'.T15.(/.234) C C-14 .//./.8) WRITE (6.125) TERM6.T55.'='.AND.K4=1.T2.5X) ) 804 CONTINUE WRITE(6.E15.E15.TERM7.100?) 1007 FORMAT(1H1.'HEAT FLOW THROUGH THE ROOM SURFACES(BTU/HR)'.T5.12.TERM6'.913) Q6 F0RMAT(//. IR. &'OVERALL ROOM HEAT LOSS = SUM OF U*A*(X(8)-T0UT)'. & T65.123) C WRITE (6.914) Q7 FORMAT(//.'Q8 = '.QCDT 803 F0RMAT(1X//. & T65.T45.802) 802 FORMAT(IX.OVERALL ROOM HEAT LOSS = SUM OF U*A*(75-T0UT)'./.fQRT'.E19. & //.8.T5.8) ) WRITE(6.Il.EQ.915) Q8 FORMAT(//.E15. &'=*.QRT. & //.'BASED ON TOTAL AREAS.//.T5.T55..912) Q5 F0RMAT(//.ALHS 125 F0RMAT(1X///. & T28.'Q7 = '.8) C WRITE(6.T18 'A POSITIVE VALUE DENOTES A LOSS FROM THE SURFACE'. & T65.8) WRITE (6.8) WRITE (6.I1.T5.12.803) IR.123) WRITE(6.QCD(I)'.TERM7'. &//.T48./) WRITE (6.123) "' WRITE(6. &'= CONDUCTION THROUGH THE CLOTHING') C NEW PAGE WRITE (6. &'ALHS (SHOULD BE = TERM6+TERM7)'.T5.'RADIATION EXCHANGE BY PERSON.8) WRITE (6.T35.QR(K4).T5.'PER UNIT AREA HEAT FLOW '.'Q5 = '. &*STD.E19.T5.E15.T70. &T65.8) WRITE (6.123) C PRINTING THE INPUT DATA AS A CHECK AT THE END OF THE PROGRAM WRITE (6.6) 25 FORMAT(1H1.6 QRT = QR(IR)*RAREA(IR) QCVT = QCV(IR)*RAREA(IR) QCDT = QCD(IR)*RAREA(IR) IF(CP .'QCDT') DO 804 IR = 1. &'SUM OF NET OUTWARD RADIATION FROM THE SURFACES'.QCD(K4).CONVECTIVE EXCHANGE.'ECHO OF INPUT DATA (AS A CHECK)'.'='.QCVT.E15.T20.T45.T2.T45.'='.3(E15. &'Q4= TOTAL HEAT LOST FROM SURFACES TO AIR BY CONVECTION1.801 912 913 914 915 & T75.'QCVT*.T5.'=*././/.. &//.6) THEN QRT = QR(IR)*PRAREA QCVT = QCV(IR)*PRAREA QCDT = QCD(IR)*PRAREA ENDIF C WRITE(6.T8.*QCV(I)'.QCV(K4).'='.T15.'Q6 = '.T35.'.'=\E15.E15. &'HEAT LOST THROUGH SURFACES TO THE OUTSIDE BY CONDUCTION'.12.25) (K4.T18.8.8. 1)AM1.'ASHRAE DESIGN HEAT LOSS'. 12.F6.'= '.'NSIG = '..'FCL ='.1) WRITE(6.1521) PAREA F0RMAT(1X//. &'PA (MM HG) = '.24)AICL.6) WRITE(6.1) C-15 . &'HC ='. &//T5.1523) Q6 +QACT3 F0RMAT(1X//.HC FORMAT (1X.8.'SENSIBLE HEAT (PEOPLE) BTU/HR/PERSON ='.T50.F5.123) C C 31 24 4 7 WRITE (6./. F10.ITMAX FORMAT(IX.1520) FORMAT( 1H1.T50.PA FORMAT(1X/.E15.'N = '.K3=1. &//T5.460.EQ.D0 FORMAT(T25.'MAX.HT WRITE (6.1.12X.21) (U(K3).(SQ.T5.7)Q2P/P ENDIF FORMAT (IX.' SQ.10X.3)V WRITE (6.2. WRITE(6.'NUMBER OF EQUATIONS* .O. NO.66)T0UT .1003)HREF.1) WRITE (6.//./.'T0TAL PANEL AREA '.C') WRITE (6.921) RATIO IF( P.S HAVE BEEN CALCULATED BY THE PROGRAM.'= '.K3=1.F5.K3=1.2501)XPREF-460.F5.23) (CI(K3).10X.'ACTUAL DESIGN HEAT LOSS'.E15.T50 .6)CFM.T5 'AICL ='.6) WRITE(6./) IF(CP) THEN C C 2501 C 1521 C WRITE(6.8) WRITE (6.7)Q2P ELSE WRITE (6.22) (HI(K3).' CLO'.*PANEL TEMPERATURE'.2.6) THE CI.8) WRITE (6.BTH.211)(K3. ITMAX = '.M).SLOPE C C WRITE(6.T5. T35.14) WRITE(6.NSIG.'NUMBER OF SIGNIFICANT DIGITS' T35.EFF WRITE (6. T35 .4)RH. RH = '.'RELATIVE HUMIDITY.K3=1.T5.K3=1.K3=1.WRITE(6.T5.'= '.6) C C C 1520 C C C C C C C FOR THIS 1 CASE WRITE(6.T50.1522) Q7 + QSTD3 FORMAT(1X//.T5.T5.78)ALTH.6) WRITE(6.235) (EPSI(K3).F10.FCL.12.609) (FP(K3).3.F10.5)P WRITE(6.8)Q2L WRITE (6. <KCAL/HR.ACH WRITE (6.DO WRITE (6.E15.6) WRITE(6. OF ITERATIONS'.DO) THEN WRITE(6.'= '.FT ') ENDIF C C C 1522 C 1523 WRITE(6.31)N. 1541) AUST-460.F10.1) WRITE(6.1539) TOF 1539 FORMATQX//.'= \F5.1538) X(9)-460.*= '.'PDIFF4'.2.D0*QRP/(Q1+Q2) FORMAT(1X//.1535) PARM3 1535 FORMAT(1X//.T50.'HEAT OUTPUT PER UNIT AREA'.1 ) ENDIF C C WRITE(6.1529) Ql +Q2 F0RMAT(1X//.F5.T50.'ROOM AIR TEMPERATURE'.'PDIFF3'. EFF.1532) PAREA*100.'PARAMETER l'.1540) ERFFPS 1540 FORMAT(1X//.T5.1 ) C WRITE(6.T50.6) C ENDIF C IF(HF) THEN CONTINUE ELSE C WRITE(6.'FLOOR TEMPERATURE'.* SUPPLY AIR TEMPERATURE'.'= '.T5.1526) PDIFF2 FORMAT(1X//.' % ') C WRITE(6.25) C ASDF 0UT(IVAR.F10.T5.2.T60.1537) X(8)-460.'= '.T5.F5.T5.= *.2 ) C WRITE(6.F5.U.' %'.1525) Q5 +QACT3 FORMAT(1X//.1 ) C C IF(CP) THEN CONTINUE ELSE C WRITE(6. RADIANT FIELD'.'.T50.1) = ACH C-16 .*= '.F7.2.1536) X(5)-460.' %') WRITE(6.D0/(RAREA(6)) FORMAT(1X//.F9.T5.1542) X(10)-460.1531)100.'PERCENT CEILING COVERED BY PANELS'.T5.T50.T5.F7.T5.'ACTUAL HEAT INPUT '.F7.'= \F7.D0 1537 F0RMAT(1X//.T5.1 ) C WRITE(6.D0 1542 F0RMATC1X//.'= '.'= '.F10.1534) PARM1 1534 FORMAT(1X//.*PARAMETER 3'.1533)Q1/PAREA 1533 FORMAT(1X//.1) WRITE(6.DESIGN HEAT LOSS 2'.T50.' %*.= '.'= '.' = '.'= '.'PDIFFl'.*PERCENTAGE RADIATION './) WRITE(6.'OPERATIVE TEMPERATURE' ..D0 1538 FORMAT(1X//.T50.*MEAN RADIANT TEMPERATURE'.T5.'A./) WRITE(6.2.T50./) IF(CP) THEN C C 1531 C 1532 WRITE(6.2.T5.'C0NDUC.T5.1 ) ENDIF C C TRYING TO WRITE IN THE DESIRED FORMAT C 0UT(25.T60.F5.T50.S.T60.T5.5) C WRITE(6.T5.1524) PDIFF1 FORMAT(lX//.F7.2.1 ) C WRITE(6.1) WRITE(6.' %'J) WRITE(6.F10.'C0NDUC.'= *.DESIGN HEAT LOSS l'.D0 1536 FORMATQX//.T.'= '.T5.'= '. * = *.F5.1527) Q5 +Q3 FORMAT(1X//.T5.'PDIFF2'. '= ' .T5.T5.F5.'= *.T50.T50.C 1524 C 1525 C 1526 C 1527 C 1528 C 1529 C 1530 C WRITE(6.T50.F10..T50.1528) PDIFF3 FORMAT(1X//.F7.' %'.1 ) C WRITE(6.T60. & T60.1530) PDIFF4 FORMAT(1X//.D0 1541 F0RMAT(1X//.T50. 2) PAREA 0UT(IVAR.10) = Ql +Q2 OUT(IVAR. SQ FT POUTC20 'A. F C-17 . F POUTC16 'ROOM AIR TEMPERATURE. BTU/HR POUTC 'PERCENTAGE DIFFERENCE 3 POUTCIO 'ACTUAL HEAT INPUT.19)= ERFFPS OUT(IVAR.DO/(RAREA(6)) IF(CP) THEN 0UT(IVAR.7) PDIFF2 OUT(IVAR. DEG.14)= Ql/PAREA ENDIF OUT(IVAR.16)= X(8)-460.15)= X(5)-460.4) Q6 +QACT3 0UT(IVAR.3) Q7 + QSTD3 0UT(IVAR.17)= X(9)-460.18)= TOF OUT(IVAR.5) PDIFF1 0UT(IVAR. SQ FT POUT( 'ASHRAE DESIGN HEAT LOSS. BTU/HR.D0 C C C OUTl(IVAR. DEG. BTU/HR POUT( 'PERCENTAGE DIFFERENCE 1 POUTC 'CONDUCTION DESIGN HEAT LOSS 1. DEG. BTU/HR POUTC11 'PERCENTAGE DIFFERENCE 4 POUTC12 'PERCENTAGE RADIATION POUTC13 'PERCENT CEILING COVERED BY PANELS POUTC14 'HEAT OUTPUT PER UNIT PANEL AREA.4)= PARM1 PARM3 ALTH BTH C C C THIS IS THE END OF THE IVAR=1. DEG. BTU/HR POUT( 'ACTUAL DESIGN HEAT LOSS.SQ FT POUTC15 'FLOOR TEMPERATURE.D0 OUT(IVAR.D0 OUTCIVAR.9) PDIFF3 OUT(IVAR.21)= X(10)-460.6) Q5 +QACT3 0UTCIVAR.8) Q5 +Q3 OUT(IVAR.ll)= PDIFF4 OUT(IVAR.DO*QRP/(Q1+Q2) OUT(IVAR113)= PAREA*100.D0 C C 0UT(IVAR. BTU/HR POUTC 'PERCENTAGE DIFFERENCE 2 POUTC 'CONDUCTION DESIGN HEAT LOSS 2. DEG.23) .12)= 100. C C ASDF FIRST LINE POUT( 'INFILTRATION AC/H POUTC 'PANEL AREA REQUIRED .23) ETC HAVE CONSECUTIVE VALUES OF X(10)-460.0UT(IVAR.NVAR LOOP C 1550 CONTINUE C C OUT(l.0UT(2.S.D0 OUT(IVAR.T.20)= AUST-460.3)= 0UT1(IVAR. F POUTC18 'OPERATIVE TEMPERATURE.U. BTU/HR. F POUTC19 'EFFECTIVE RADIANT FIELD.l)= OUTl(IVAR. F POUTC17 'MEAN RADIANT TEMPERATURE.2)= OUTl(IVAR. ALL GLASS 'ONE WALL.7(F8.I).7(F8. HALF GLASS 'ONE WALL. ALL GLASS 'PANEL TEMPERATURE.2500) TITLE FORMAT(T1.A45.7(F8.T2.T50.1X) ) 3111 CONTINUE C DO 3503 I = 1.3502) POUTl(I).2 WRITE(6.J=1.T50.11 WRITE (6.3511) POUT(l).T2.NVAR) 3502 F0RMAT(/.T50.1.'. ALL GLASS-SECOND WALL. POUTl(I+2) C-18 . HALF GLASS 'TWO WALLS.1X) ) c c WRITE (6.P0UT(21) = 'SUPPLY AIR TEMPERATURE.5001) POUT(I).1X) ) C C THE FOLLOWING IS FOR PANEL HEATING C IF (CP) THEN DO 3111 I = 2. (0UT1(J.F 'PARAMETER 3.T50.1.1X) ) C 3550 CONTINUE ENDIF C C THE FOLLOWING IS FOR THE GLASS CASE C DO 3521 I = 1.J=1.T50.I).1.NVAR) FORMAT(/.NVAR) 5001 F0RMAT(/.1X) ) 5000 CONTINUE DO 3550 I = 15.T50.T2.T2.A45./) WRITE (6.J=l.T2.A50. F "U-FACTOR CASE c c c C C C C C = = = = = = = = = REWRITING ONLY THE REQUIRED OUTPUT WRITE(6.SQ FT.1520) ABOVE GIVES NEW PAGE 2500 C C 3523 C WRITE(6.3522) I. DEG.J=1.A45. (J.(OUT(J.(OUT(J.NVAR) 3120 FORMAT(/.A45.A45.7(F8. DEG. BTU/HR.CE *.A45. (0UT(J.3120) POUT(I).3551) POUT(I).I).NVAR) 3551 FORMAT(/.7(I8.21 WRITE (6.4.NVAR) 3511 FORMAT(/.J=1.1X) ) 3503 CONTINUE ELSE C C C THE FOLLOWING IS TO BE USED FOR CONVECTIVE HEATING C DO 5000 I = 3.1.(OUT(J.5 C WRITE(6. DIMENSIONLESS 'NO GLASS IN ANY WALL 'ONE WALL.7(F8.T2.20 WRITE (6.J=1. F POUTl(l) P0UT1(2) P0UT1(3) P0UT1(4) POUTl(5) POUTl(6) POUTl(7) POUTl(8) POUTl(9) 'PARAMETER 1.I).1).3523) P0UT1(9). F(10).3551) POUT(I).'.0-Z) LOGICAL CP.NVAR) 4002 CONTINUE C ELSE C C DO 5002 I = 3.'.J=1.I).US l') WRITE(8.NVAR) 4000 CONTINUE C C THE FOLLOWING IS FOR CONVECTIVE HEATING DO 4002 I = 1.NVAR) 5003 CONTINUE ENDIF C C 999 STOP END C C C C C******* SUBROUTINE STARTS *****fr&****ftifo^^^ C SUBROUTINE FCN(X.4012) WRITE(8.'.I).3502) POUTl(I).': *.COOL.N. A45) 3521 CONTINUE C C C WRITING WITH DEVICE TYPE 8 SO AS TO GET AN OUTPUT IN A DIFFERENT FIL C C WRITE(8.I).'CASE NUMBER '.12.'.4010) WRITE(8.PN OFF') 4012 FORMAT(Tl.RF CANCEL') 4010 FORMAT(T1.(OUT(J.4014) WRITE(8.I).l).F.NVAR) 5002 CONTINUE DO 5003 I = 15.11 WRITE (8.J=1.J=1.'.3120) POUT(I).(0UT(J.3511) POUT(l).T2.PAR(10) COMMON /OUT/ TOUT C-19 .J=l.J=1.2500) TITLE C C C WRITE (8.PAR) IMPLICIT REAL*8 (A-H.(OUT(J.4011) WRITE(8. (0UT1(J.HF DIMENSION X(10).3522 F0RMAT(/.2 WRITE(8. (OUT(J.FO OFF') 4013 FORMAT(Tl.NVAR) C C IF(CP) THEN C C THE FOLLOWING IS FOR PANEL HEATING DO 4000 I = 2.LL 120') 4011 FORMAT(Tl.5001) POUT(I).20 WRITE (8.4013) 4014 FORMAT(T1.21 WRITE (8. C C C C C C C C C C COMMON /CEL/ XCEL(ll) COMMON /COMF/ AM1.V.YCP(25).HC.BCP(25) FSURF APPEARS IN MPROG.1714D-8 C C IF(CP) THEN C FOR SURFACES 1 TO 6 (CEILING IS NOW OF AREA PRAREA) .Q5.S TO CELSIUS VALUES DO 114 11= 1.5) IS USED AND NOT FS(5.UQ2P.8D0 C I II I IIIIIIl II I11 II III IF(CP) THEN XPANEL = X(10) ENDIF C C C C C C C C C C C C IM IIII I IIIH IIIH I TOTAL AREA OF PANELS.COOL.FSCP MATRIX C IS USED INSTEAD OF FS.ALHS COMMON /DIMEN/ ALTH.5)* AREA(l) = FS(5.Q7.Q2L COMMON /Q/ Q1.TERM7.N 114 XCEL(Il) = (X(I1) -492D0 )/1.6 QROUT = EPSI(J6) *SIGMA* (X(J6)**4) QRIN =O.1) BECAUSE FS(1.QCD(6) COMMON /U/ U(6).PAREA COMMON /PAN3/ NCP COMMON /PAN4/ FSCP(25.QRP COMMON /QI/ QR(6).HIP.1)* AREA(5) AND EVERYTHING HERE IS PRORATED TO THE AREA OF THE SURFACE EG.HIDOWN COMMON /FSURF/ FS(6.QACT3.PRAREA ARE OBTAINED FROM THE MAIN PROGRAM THROUGH THE COMMON STATEMENT.HI(6).HT COMMON /AREAS/ RAREA(6) COMMON /PAN1/ CP.Q2P. AREA(l) SIGMA = 0.RH COMMON /AIR1/ P.EPSIP.EFF.EPSI(6) COMMON /UP/ UP.QNET3.HF COMMON /PAN2/ PRAREA.SLOPE COMMON /TERM/ TERM6.AICFM. DO 111 J6 = 1.ACH.6) COMMON /FPEOP/ FP(6) COMMON /GRAD/ HREF.QCV(6). AND FCN PAN3 APPEARS IN MPROG.Q3.QP5.6 QRIN = QRIN + EPSI(J7)* SIGMA* FSCP(J6..?A.AICL.J7)* ( X(J7)**4) C-20 .PAREA AND REST OF CEILING AREA .Q8 COMMON /QNET/ QNETl. AND FCN' PAN2 APPEARS IN MPROG.XMULT COMMON /CONV/ HIUP.Q6. DO 113 J7 = 1. AND FCN PAN1 APPEARS IN MPROG.QNET2.BTH. FS(1.ALCP(25). AND FCN CONVERTING RANKINE TEMP. CALCULATING THE RADIATIVE HEAT TRANSFER NET-GOING-OUT .UQ2L.OUA COMMON /QP/ QCVP.FCL.CIP.CI(6).DO C FROM OTHER WALLS.Q4.CFM.QNETP2. FOR EG.25) COMMON /DIMP/ XCP(25). AND FCN PAN4 APPEARS IN MPROG..Q2.QSTD3. TO MAKE THE PANELS RADIATE ONLY TO THE C FLOOR.712D0 . IT SHOULD BE TAKEN INTO ACCOUNT WHEN THE ULTIMATE C SOURCE OF HEAT IS EXAMINED.LT.NCP+10 C******** INFRARED STARTS********* DO 6002 JJ = 1 .' C C CONVECTION COEFFICIENTS C C FOR USE INSIDE THE PROGRAM. IF NO PANELS DO 603 J6 = 1. WHEN THE HEAT FLOW IS VERTICALLY C DOWNWARD .X(8) ) THEN HI(5) = HIDOWN C-21 .K)= O.162D0 .K)* (XPANEL**4) 808 CONTINUE QR(J6) = QROUT -QRIN.STANDARD ASHRAE VALUE C HIUP IS FOR HEAT FLOW VERTICALLY UPWARD.I 113 CONTINUE C FROM THE PANELS. RADIATION TO THE WALLS IS MADE EQUAL TO ZERO C C IF THIS IS NOT DESIRED. STANDARD ASHRAE VALUE C HIDOWN IS THE CONVECTION COEFF. MAKE STATEMENTS FROM "INFRARED STARTS" TO C "INFRARED ENDS" AS COMMENT LINES.. Ill CONTINUE C C NOTE: THE HEAT BALANCE OF THE PORTION COVERED BY PANELS IS NOT C CONSIDERED HERE . C C ELSE C I. C C C DO 808 K = 11.6 QROUT = EPSI(J6) *SIGMA* (X(J6)**4) QRIN =0.E.K) = ( ALCP(K)*BCP(K) )/(ALTH*BTH) C******** INFRARED ENDS ********** QRIN = QRIN + EPSIP*SIGMA* FSCP(J6.DO 6002 CONTINUE FSCP(5. HI(5) & HI(6) MUST BE CHANGED TO C VALUES NOT INCLUDING RADIATION C HIDOWN = 0. 4 FSCP(JJ.J7)* ( X(J7)**4 ) 604 CONTINUE QR(J6) = QROUT -QRIN 603 CONTINUE C ENDIF C C NOW COMPUTING THE CONVECTIVE AND CONDUCTIVE PARTS OF HEAT TRANSFER C NET-GOING-OUT AND HENCE MAKING A HEAT BALANCE .6 QRIN = QRIN + EPSI(J7)* SIGMA* FS(J6. C C FOR THE CONVECTIVE CASE C — — C FOR THE FLOOR IF( X(5).D0 DO 604 31 = 1. C HIUP = 0. ^ ^ ^ ^ ^ ^ ^ ^ ^ C FOR A SPECIAL CASE . .4 CI I IMIIHI(IW) = 0..26D0* (DABS(X(IW)-X(8) )**0. C C THE EQUIVALENT DIAMETER DE FOR THE CEILING OR THE FLOOR IS.X(8) ) THEN HI(5) = HIDOWN ELSE HI(5) = HIUP ENDIF C C C WALLS DO 950 IW = 1.LT.XCEL(K2))/HEI) **0.X(8) ) THEN HI(6) = HIUP ELSE HI(6) = HIDOWN ENDIF C C FOR THE WALLS C C HI BASED ON DELTA T . H IN METRES.BAUMAN*S CORRELATION 1 FT = 0. IF( X(6).3 AND 4. IF THE PANELS ARE IN THE CEILING.3048 M (EXACT) C FOR WALLS 1. DIFF.2.68 TO GET BTU/HR. IF( X(6).05D0) 950 CONTINUE C C THE ENDIF FOR THE IF (CP) FOLLOWS.3048DO HIBAU= 2.. C C UNHEATED CEILING PORTION .SQ FT.4 HEI = HT*0.. C NOTE : THIS CORRELATION NEEDS TEMP.03D0*( DABS((XCEL(8).32D0 )/(HT**0.D0*(ALTH+BTH) ) C C HEATED CEILING CASE C IF(CP) THEN C C THAT IS..25D0) C C TRYING ASH STD FOR THE FLOOR C IF( X(5).LT.F C DO 52 K2=l. C-22 .C ( DIV BY 5.ELSE HI(5) = HIUP ENDIF C FOR THE CEILING. IN CELSIUS.68D0 52 CONTINUE C C NOW TRYING THE CORRELATIONS GIVEN BY MIN AND SCHUTRUM .. C X(5)= FLOOR AND X(6) = CEILING TEMPS. DE = 4.X(8) ) THEN HI(6) = HIUP ELSE HI(6) = HIDOWN ENDIF C FLOOR C HI(5) = 0.. C FOR HEATED CEILING PANEL OR HEATED FLOOR HIP IS GIVEN JUST BEFORE USE.25D0)/(DE**0.LT.32D0 ) HI(IW) = 0.22D0 ) HI(K2) = HIBAU/5. CAND THE RESULTING HI IS IN W/M**2.DO* RAREA(5)/ (2.041D0*( DABS( X(8)-X(5) )**0.29D0*( DABS(X(IW)-X(8) )**0. (X(8)-H*G) ) QCD(5)= CI(5)*(X(5)-T0UT) F(5)= QR(5) + QCV(5) +QCD(5) C IF(CP) THEN C I.X(8) ) THEN HI(6) = HIDOWN ELSE HI(6) = HIUP ENDIF C C WALLS " DO 1200 IW = 1.39* (DABS(X(8)-X(5) )**0.X(8) ) THEN HI(5) = HIUP ELSE HI(5) = HIDOWN ENDIF C UNHEATED FLOOR PORTION IF( X(6).05D0) 1200 CONTINUE C C THE ENDIF FOR THE IF (HF) FOLLOWS.29D0*( DABS(X(IW)-X(8) )**0.LT.LT. ABOVE ARE TO BE CHANGED C X(5) IS NOW THE CEILING TEMP & X(6) IS THAT OF THE FLOOR. ENDIF C C QCV(1)= HI(1)*( X(l)-X(8) ) QCD(1)= CI(1)*(X(1)-T0UT) F(l)= QR(1) + QCV(l) +QCD(1) C QCV(2)= HI(2)*(X(2)-X(8)) QCD(2)= CI(2)*(X(2)-T0UT) F(2)= QR(2) + QCV(2) +QCD(2) C QCV(3)= HI(3)* (X(3)-X(8)) QCD(3)= CI(3)*(X(3)-T0UT) F(3)= QR(3) + QCV(3) +QCD(3) C QCV(4)= HI(4)* (X(4)-X(8)) QCD(4)= CI(4)*(X(4)-T0UT) F(4)= QR(4) + QCV(4) +QCD(4) C C QCV(5)= HI(5)* ( X(5) -X(8)) C QCV(5)= HI(5)* ( X(5) . C IF(HF) THEN C C CEILING C***********HI(5) = 0. IF A HEATED FLOOR IS USED .31DO )/(DE**0.26D0* (DABS(X(IW)-X(8) )**0.32D0 )/(HT**0.4 CI I I I I IIHI(IW) = 0.32D0 ) HI(IW) = 0.ENDIF C C HEATED FLOOR C C THAT IS.. C QCV(6) IS NOW PER UNIT PRAREA C-23 .E.08) IF( X(5). WITH CEILING PANELS. DO.1289706D-4 C12=-0. =+ QCV(6)= HI(6)*( X(6) -( X(8)+(8.35D0* ( 43.1..2478068D-8 C13=6.DO-PA) TERM5= 0. 1.D0)**4 . IF(X(8).D0-EFF) TERM2= 0.E-50)THEN ALNPWS = O. COEFF.02700133 Cll=0.4 C9=-ll. BASED ON TEMP. OF CLOTHING BASED ON VELOCITY.715 CONVERTS PSI TO MM HG C8=-10440.XCEL(8) ) C ALHS REPRESENTS THE NET CONDUCTION THROUGH THE CLOTHING ALHS = TERM1 .4D-8*FCL*( (XCEL(7) +273.0023* AMI*(44.E.LT.DO ELSE ALNPWS=C8/X(8) +C9 +C10*X(8) +C11*(X(8)**2) +C12*(X(8)**3)+ & C13*DL0G(X(8)) ENDIF PA=DEXP(ALNPWS) * 51. ARE CONVERTED TO CELSIUS) C THE FOLLOWING IS TO EVALUATE PA AT THE CURRENT AIR TEMP(R): X(8) C 51.D0-H)*G ) + ) QCD6 = CI(6)*(X(6)-TOUT) IF ( PRAREA . THE GREATER OF THE TWO VALUES IS USED HCV = 10. PRAREA END IF ELSE C I. (FORCED CONV) C HCTD.\ C** C C** C QCV(6) = HI(6)* ( X(6)-X(8) ) .TERM2 -TERM3 -TERM4 -TERM5 TERM6= 3.4D0* DSQRT(V) C-24 .DO) TERM4= 0. DIFF.LT.(XCEL(9)+273.5459673 C TO TAKE CARE OF NEGATIVE X(8)!M (DURING ITERATIONS).D0-H)*G ) ) QCD(6)= CI(6)*(X(6)-TOUT) F(6)= QR(6) + QCV(6) +QCD(6) ENDIF C C C C C C THE COMFORT EQUATION ( TEMPS.DO ELSE QCD(6) = QCD6 F(6)= QR(6) + QCV(6) +QCD(6) C NOTE:THIS HEAT BALANCE IS PER UNIT OF REDUCED CEILING.061D0*AMl*(l..D0-EFF) -50. C OBSERVE THE IF LOOP. FOR FREE -CONV.0D-4) THEN QCD(6) = O.DO-EFF)-PA ) TERM3= 0. QCV(6)= HI(6)*( X(6) -X(8) ) C QCV(6)= HI(6)*( X(6) -( X(8)+(8. WITHOUT ANY CEILING PANELS.D0)**4 ) C C HCV IS THE CONV.D0-0.42D0*( AM1*(1.0014DO*AM1*(34.DO X(6) = XPANEL F(6) = O.715D0* RH C C TERM1= AM1*(1.29466692 C10=-0. ( X(9)**4 ) ENDIF C C C TO GET THE HEAT INPUT TO THE ROOM BY THE PANELS BY C CONVECTION AND RADIATION.( X(9)**4 ) ELSE FPTOT = O.25D0 ) HC= HCTD IF (HCV.DO DO 607 IP = 1.HCTD= 2.05*( (DABS(XCEL(7)-XCEL(8)) )**0.7D0.25D0 )/(DE**0. IF (CP) THEN C THIS DIVISION OF FP(6) IS QUITE APPROXIMATE.31D0 ) C •: HIP = HIUP ENDIF C IF(CP) THEN UQPOUT = EPSIP *SIGMA * (XPANEL**4) QPOUT = UQPOUT*PAREA C RADIATION OUTWARDS FROM ALL THE PANELS QPIN =0.08D0) C C*****HIP = 0.D0 C-25 .GT. FP6 = FP(6)*PRAREA/ (ALTH*BTH) FPP = FP(6)*PAREA / (ALTH*BTH) FPTOT = O.DO DO 608 IP = 1.31D0* ( DABS(X(8)-X(6) )**0. TO INCLUDE IN THE HEAT BALANCE OF THE ROOM IF(CP) THEN C FOR THE PANELS.18*AICL*ALHS F(8)= RHS.IN THE FLOOR C HIP = 0.ALHS C .DO-EFF) C F(7)= 35.25D0) C ENDIF C IF(HF) THEN C FOR THE PANELS.041D0*(DABS(X(8)-XPANEL)**0.TERM8 -0.5 FPTOT = FPTOT + FP(IP)* (X(IP)**4) 607 CONTINUE FPTOT = FPTOT + FP6*(X(6)**4) + FPP*(XPANEL**4) F(9) = FPTOT .HC) THEN HC= HCV ENDIF CC CC CC CC TERM7= FCL*HC*( XCEL(7)-XCEL(8) ) RHS = TERM6 +TERM7 TERM8= 0.31D0 )/(DE**0.D0*(ALTH+BTH) ) C HIP = O.6 FPTOT = FPTOT + FP(IP)* (X(IP)**4) 608 CONTINUE F(9) = FPTOT .XCEL(7) CALCULATION OF MEAN RADIANT TEMPERATURE (NEGLECTING REFLECTIONS).39* (DABS(X(8)-XPANEL)**0. DE = 4.032DO*AM1*(1.IN THE CEILING C THE EQUIVALENT DIAMETER DE FOR THE CEILING OR THE FLOOR IS.DO* RAREA(5)/ (2. 31D0* ( DABS(X(8)-X(6) )**0.DO* RAREA(5)/ (2.D0 DO 607 IP = 1.31D0 )/(DE**0.08D0) C C*****HIP = 0.D0 C-25 . TO INCLUDE IN THE HEAT BALANCE OF THE ROOM IF(CP) THEN C FOR THE PANELS.IN THE FLOOR C HIP = 0.IN THE CEILING C THE EQUIVALENT DIAMETER DE FOR THE CEILING OR THE FLOOR IS.HCTD= 2.18*AICL*ALHS .TERM8 -0.39* (DABS(X(8)-XPANEL)**0.D0 DO 608 IP = 1.( X(9)**4 ) ENDIF C C C TO GET THE HEAT INPUT TO THE ROOM BY THE PANELS BY C CONVECTION AND RADIATION.05*( (DABS(XCEL(7)-XCEL(8)) )**0.GT.D0-EFF) C F(7)= 35. FP6 = FP(6)*PRAREA/ (ALTH*BTH) FPP = FP(6)*PAREA / (ALTH*BTH) FPTOT = 0.D0*(ALTH+BTH) ) C HIP = 0.041D0*(DABS(X(8)-XPANEL)**0.25D0) C ENDIF C IF(HF) THEN C FOR THE PANELS.25D0 )/(DE**0.( X(9)**4 ) ELSE FPTOT = 0. IF (CP) THEN C THIS DIVISION OF FP(6) IS QUITE APPROXIMATE.31D0 ) C = = HIP = HIUP ENDIF C IF(CP) THEN UQPOUT = EPSIP *SIGMA * (XPANEL**4) QPOUT = UQPOUT*PAREA C RADIATION OUTWARDS FROM ALL THE PANELS QPIN =0.6 FPTOT = FPTOT + FP(IP)* (X(IP)**4) 608 CONTINUE F(9) = FPTOT .7D0. DE = 4.HC) THEN HC= HCV ENDIF CC CC CC CC TERM7= FCL*HC*( XCEL(7)-XCEL(8) ) RHS = TERM6 +TERM7 TERM8= 0.25D0 ) HC= HCTD IF (HCV.5 FPTOT = FPTOT + FP(IP)* (X(IP)**4) 607 CONTINUE FPTOT = FPTOT + FP6*(X(6)**4) + FPP*(XPANEL**4) F(9) = FPTOT .032D0*AM1*(1.XCEL(7) F(8)= RHS.ALHS ___tJ_1_^_1_tJ_^J_1_1_UJ_^ C CALCULATION OF MEAN RADIANT TEMPERATURE (NEGLECTING REFLECTIONS). J7)* ( X(J7)**4 ) C ABOVE : RADIATION INWARDS FOR THIS ONE PANEL QPIN = QPIN + UQPIN* ALCP(JP)*BCP(JP) 602 CONTINUE QRP = QPOUT .8D0*3.08D0*CFM*ALTH*BTH*( X(10)-X(8) ) C Q1=NET HEAT INPUT BY AIR ENDIF „.E. CFM AICFM = ACH* (ALTH*BTH*HT)/60.6001 DO 602 JP = ll.USING A READ-IN VALUE C TO USE THE CONDUCTION THROUGH THE CLOTHING FOR UQ2P C BTU/HR = (#)* KCAL/HR.SQM * (DUBOIS AREA = 1.DO CONTINUE FSCP(JP.D0+75.973D0 C CCCCC Q2L= P*UQ2L Q2= Q2P + Q2L C Q2=HEAT INPUT BY PEOPLE AND LIGHTS C ACH= NO.D0 -TOUT ) QACT3= 1.6 INFRARED STARTS********* DO 6001 JI = 1.JI) = O.I.08D0*AICFM *( X(8) -TOUT ) XINF = X(8) + (HT-HREF)*SLOPE Q3= 1.08D0*AICFM *( 460.D0 DO 601 J7 = 1.D0 QSTD3= 1.8)SQM* 3.5) = l.4 FSCP(JP.8 BTU/HR/PERSON (STD. OF AIR CHANGES PER HOUR.=250) Q2P = P*ALHS*1.DO C******** INFRARED ENDS ************ 601 UQPIN = UQPIN + EPSI(J7)* SIGMA* FSCP(JP.08D0* AICFM* (XINF-TOUT ) C Q3=HEAT LOSS DUE TO INFILTRATION AIR C C QP4= ALTH*HT*(QCV(1) +QCV(3)) + BTH*HT*(QCV(2) +QCV(4)) & + ALTH*BTH*QCV(5) + PRAREA*QCV(6) QP5= ALTH*HT*( QCD(l) +QCD(3) ) + BTH*HT*( QCD(2) +QCD(4) ) & + ALTH*BTH*( QCD(5) ) + PRAREA*QCD(6) QP8= ALTH*HT*( QR(1) +QR(3") ) + BTH*HT*( QR(2) +QR(4) ) & + ALTH*BTH* QR(5) +PRAREA*QR(6) POUA = ALTH*HT*( U(l)+U(3) ) +BTH*HT*( U(2)+U(4) ) & + ALTH*BTH* U(5)+ PRAREA*U(6) C C C AND WHEN THERE ARE NO CEILING PANELS. Q4= ALTH*HT*(QCV(1) +QCV(3)) + BTH*HT*(QCV(2) +QCV(4)) & + ALTH*BTH*(QCV(5) +QCV(6)) Q5= ALTH*HT*( QCD(l) +QCD(3) ) + BTH*HT*( QCD(2) +QCD(4) ) & + ALTH*BTH*( QCD(5) +QCD(6) ) Q8= ALTH*HT*( QR(1) +QR(3) ) + BTH*HT*( QR(2) +QR(4) ) C-26 . AICFM = INF. MET =1. THIS IS 252.QPIN C ABOVE : NET OUTWARD RADIATION FROM ALL PANELS (BTU/HR) QCVP = ( XMULT*HIP*(XPANEL -X(8)) ) * PAREA C ABOVE : CONVECTION FROM TOTAL PANEL AREA TO AIR Ql = QCVP + QRP C Q1=NET HEAT INPUT BY PANELS (CONDUCTION NOT CONSIDERED) ELSE Ql= 1.973D0 BTU/KCAL C FOR AM1= 50. C C c C***** Q2p = p*UQ2P .NCP+10 UQPIN =0. 'BREADTH = '.WHEN FLOW IS FROM SURFACE TO AIR.1.D0/ .1.5X. STANDARD CORRECTIVE ACTION TO BE APPLIED C *** RANGE OF ERROR NUMBERS UPTO THIS NUMBER<OMITTED HERE) C502 FORMAT(1X/. B/4*0.PERPENDICULAR TO EACH OTHER & FLOOR.1. HT C IMPLICIT REAL*8 (A-H. 'HEIGHT = '.'LENGTH = '.D0+75. BTH.T5.2. C & SX.1.D0-TOUT)* OUA Q7=STD.C C C C C C C & + ALTH*BTH*( QR(5) +QR(6) ) OUA = ALTH*HT*( U(l)+U(3j ) +BTH*HT*( U(2)+U(4) ) & + ALTH*BTH*( U(5)+U(6) ) NEGATIVE OF Q4 = HEAT LOST FROM AIR TO THE SURFACES BY CONVECTION NOTE: QCVS ARE SET UP AS POSITIVE. QNET1= Ql +Q2 -Q3 +Q4 NOTE: QNET1 = 0 = F(10) IS USED FOR CONVECTIVE HEATING Q5= HEAT LOST THROUGH THE SURFACES BY CONDUCTION TO THE OUTSIDE QNET2= Ql +Q2 -Q3 -Q5 QNETP2= Ql +Q2 -Q3 -QP5 NOTE: QNETP2 = 0 = F(10) IS USED FOR PANEL HEATING.BTH.FS) C INPUT NEEDED : ALTH. C CEILING.B(4). NO OF MESSAGES TO BE PRINTED C *** 1.1) C *** 263 ERR NO.256.D0/ C THIS IS TO CALCULATE THE SHAPE FACTORS OF THE SIX SURFACES.2) C ===== RAREA(l) = ALTH * HT RAREA(3) = RAREA(l) RAREA(2) = BTH * HT RAREA(4) = RAREA(2) RAREA(5) = ALTH * BTH C-27 .1) C CALL ERRSET(209.HT. (209 ALSO) C *** 256 UNLIMITED NO. Q6= (X(8)-TOUT)* OUA Q6=OVERALL ROOM HEAT LOSS Q7= (460.6) COMMON /AREAS/ RAREA(6) DATA A/4*0. C ******** SEE HANDOUT FOR BELOW CODING EXPLANATION C CALL ERRSET(263.F6.256. OF ERROR OCCURRENCES C *** 1.F6.FS(6. NO TRACEBACK IS TO BE PRINTED C *** i.OVERALL ROOM HEAT LOSS QNET3 = Ql +Q2 -Q3 -Q6 C IF(CP) THEN F(10)= QNETP2 ELSE F(10)= QNET1 ENDIF C RETURN END C C C*************************** C* IF(CP) THEN C* ELSE C* ENDIF SUBROUTINE SHAPE(ALTH. 4 VERTICAL WALLS.O-Z) DIMENSION A(4).F6.2.3X.?ROOM DIMENSIONS : '. FS(1.4 A(I) = O.3) = FS(IJ4.1) DO 508 IJ4 = 4.IJ2)* RAREA(1)/RAREA(IJ2) C FOR (SURFACE.1) C FOR FS (2..5) DO 505 1=1.F12) FS(1.4 A(I) = O.3) = FS(2.DO 504 B(I) = O..6 FS(IJ.5).DO 503 CONTINUE C FOR FS(1.F12) FS(1.3).5) = F12 C C C FOR FS(1..6) = FS(1.2).6 508 FS(IJ4.4 A(I) = O.2) FS(1.5) C C USING THE PRINCIPLE OF RECIPROCITY DO 507 IJ2 = 2.TO ZERO DO 503 IJ = 1..2) = F12 C C C C C FOR FS (1..B.IJ) = O.DO C-28 . DO 509 1=1.DO A(2) = ALTH B(2) = ALTH A(4) = HT B(4) = HT G = BTH C CALL SHPRL (A.DO A(2) = ALTH B(2) = ALTH A(4) = HT B(4) = BTH CALL SHPRP(A...4) = FS(1.6 507 FS (IJ2.3) FS(2.B.RAREA(6) = RAREA(5) C INITIALISING.1) = FS(1.DO C A(2) = HT B(2) = HT A(4) = ALTH B(4) = BTH CALL SHPRP(A.4 A(I) = O.B.DO 505 B(I) = O.DO 509 B(I) = O. DO 504 1=1.3) = F12 C C C SINCE 2 AND 4 ARE BOTH OF THE SAME AREA AND ORIENTATION TO 1.G.DO 506 B(I) = O.F12) FS(1. DO 506 1=1. B.6) C C C THESE ARE WRITTEN OUT AT THE END C ******** FOR PANELS IN THE CEILING.4).K) .514) (K .F12) FS(2.F12) FS(5.B. C***** PRINTING OUT RESULTS C TO COMPARE WITH THE EARLIER CASE C WRITE (6.G.4) = F12 FS(2.DO A(2) = BTH B(2) = BTH A(4) = ALTH B(4) = ALTH G = HT C CALL SHPRL (A.6) = FS(2.DO 513 B(I) = O.6) C514 FORMAT (1X/.4 A(I) = O.6).T35.2) = FS(2.K=1.T15.DO A(2) = BTH B(2) = BTH A(4) = HT B(4) = HT G = ALTH C FOR C CALL SHPRL (A.5) = F12 c c FS(2.6) C C FOR FS(5.2) DO 512 IJ4 =5.6 C WRITE (6.K=1. C & 'WITHOUT THE CEILING PANELS'.4) = FS(IJ4.6) FS(4.5) = FS(5... NCP IN NUMBER.5) USING THE PRINCIPLE OF RECIPROCITY DO 511 IJ2 = 3. DO 510 1=1.A(2) = BTH B(2j = BTH A(4) = HT B(4) = ALTH CALL SHPRP(A.(FS(I.F12) FS(2.2) C FS(3.5) FS(4S6) = FS(2.6) = F12 C FS(6.4) FS(3.4) = FS(3.DO 510 B(I) = O.9X) ) C C DO 516 1=1.4 A(I) = O. //..6) = FS(1.5) = FS(1.6 511 FS (IJ2.G.6) C C-29 . DO 513 1=1.B.'WALL-TO-WALL SHAPE FACTORS'.5) FS(3.6(I3.515)1..6 512 FS(IJ4.5) = FS(2.5X.IJ2)* RAREA(2)/RAREA(IJ2) C FOR (SURFACE. C***** PRINTING OUT RESULTS C TO COMPARE WITH THE EARLIER CASE C WRITE (6.IJ2)* RAREA(2)/RAREA(IJ2) C FOR (SURFACE.K=1.4) = F12 C FS(2.2) DO 512 IJ4 =5.4 A(I) = 0.5) C USING THE PRINCIPLE OF RECIPROCITY DO 511 IJ2 = 3.5) = FS(1.. DO 510 1=1.6 512 FS(IJ4.6) C C FOR FS(5.2) C FS(3. NCP IN NUMBER.F12) FS(2.B. DO 513 1=1.4) = FS(IJ4.4 A(I) = O. C & 'WITHOUT THE CEILING PANELS'.D0 510 B(I) = O. //.5X.6) C-29 .T35.6) = FS(2.6(I3.G.5) = F12 C C C FOR FS(2.6) = F12 C FS(6.DO 513 B(I) = O.5) = FS(2.B.6) = FS(1.4) FS(3.5) FS(3.DO A(2) = BTH B(2) = BTH A(4) = HT B(4) = HT G = ALTH C CALL SHPRL (A.T15.2) = FS(2..K) .4).6) FS(4.515)1.6) C C C THESE ARE WRITTEN OUT AT THE END C ******** FOR PANELS IN THE CEILING.6).(FS(I.514) (K .9X) ) C C DO 516 1=1..6 511 FS (IJ2.6 C WRITE (6.5) = FS(5.6) = FS(2.F12) FS(2.K=1.6) C514 FORMAT (1X/.G.4) = FS(3.'WALL-TO-WALL SHAPE FACTORS'.5) FS(4.B.DO A(2) = BTH B(2) = BTH A(4) = ALTH B(4) = ALTH G = HT C CALL SHPRL (A.A(2) = BTH B(2) = BTH A(4) = HT B(4) = ALTH CALL SHPRP(A.F12) FS(5.. (209 ALSO) C *** 256 UNLIMITED NO.BCP(25) .2.YCP(J).173) ALTHjBTH.C515 C516 C C C C C C FORMAT (1X/. OF ERROR OCCURRENCES C *** 1.ALCP(25).2X) ) CONTINUE RETURN END INPUT NEEDED: DIMENSIONS OF ROOM :ALTH.BTH.BCP(J) C WRITE(6.2.6.ALCP(J). C & 5X.'ROOM DIMENSIONS : '.13) C WRITE(6.3X.6(E10.'PANEL CENTER LOCATION :'.D0/ .2) C READ(5.*Y (ALONG BREADTH) '.1.25) COMMON /AREAS/ RAREA(6) COMMON /PAN2/ PRAREA.//) C165 CONTINUE C ===== C =================== C c c RAREA(l) = ALTH * HT RAREA(3) = RAREA(l) C-30 .'TOTAL NUMBER OF CEILING PANELS = '. C & T50.2.T30.256.T5.2.1.T5. THEIR LOCATION : XCP.*)XCP(J).2.175) C175 F0RMAT(1X/.256.T5.ALCP(J).B(4).T20.F6.NCP+10 C READ(5. 'HEIGHT = '.FS(6.4.1.'LENGTH = '.F6. OF PANELS.T70.FSCP(25.YCP(J).1.F6.1) C *** 263 ERR NO.F6.YCP(25).T85..6 PART. C PANELS ARE NAMED FROM 11 ONWARDS DATA A/4*0.1) C CALL ERRSET(209.176)J.T30 'X (ALONG LENGTH) '.2.6) DIMENSION XCP(25).) NO.FCN AND SHCP C C C C ALL MATRICES FOR PANELS MUST BE DIMENSIONED FOR NCP+ 10 AT THE C LEAST FS = FSCP FOR THE FIRST 6.D0/ C ******** SEE HANDOUT FOR BELOW CODING EXPLANATION C CALL ERRSET(263.HT & FS(6.*) NCP C WRITE(6.I3.I3.T5.YCP OF CENTRE & ALCP.PAREA /PAN2/ IS COMMON TO M/PROG.BCP(J) C176 F0RMAT(1X/.F6.T50. STANDARD CORRECTIVE ACTION TO BE APPLIED C *** RANGE OF ERROR NUMBERS UPTO THIS NUMBER(OMITTED HERE) C C C C WRITE(6. NO OF MESSAGES TO BE PRINTED C *** 1.HT C173 FORMAT(1X/.'BREADTH = '.5X.174) NCP C174 FORMAT(1X/.T85.0-Z) DIMENSION A(4).XCP(J). NO TRACEBACK IS TO BE PRINTED C *** 1.BCP (LENGTH AND BREADTH) C IMPLICIT REAL*8 (A-H. B/4*0.F6.T70.F6.T15.'LENGTH*.'BREADTH') C C C DO 165 J =11. ** EACH PANEL HAS ITS CENTER LOCATED AT XCP(J).BCP(J)/2.DO C-31 .I2) = O. 205 208 DO 205 11= 1.J) DO 167 1 = 1 .DO DO 208 11= 1.DO 166 B(I) = O.T5.DO B(l) = XCP(J) .PRAREA.6 DO 208 12= 1. BECOMES PAREA = O.DO B(l) B(2) B(3) B(4) XCP(J) XCP(J) YCP(J) YCP(J) A(2) A(4) ALTH BTH G ALCP(J)/2.DO 167 B(I) = O.I2) = FS(I1. FOR FSCP(5. A(I) = O. NCP IN NUMBER.I2) C WITH CEILING PANELS THE REST OF THE CEILING AREA PRAREA.25 FSCP(I1.J) = F12 C C C FOR FSCP(1.207) PRAREA C207 FORMAT(1X/.RAREA(2) = BTH * HT RAREA(4) = RAREA(2) RAREA(5) = ALTH * BTH WITHOUT CEILING PANELS RAREA(6) = RAREA(5) WITH CEILING PANELS IS GIVEN JUST BEFORE THE LOOP DO 171 ******** FOR PANELS IN THE CEILING.D0 B(3) = YCP(J) . WALL2 AND THE CEILING.DO DO 178 K = 11.YCP(J) AND HAS DIMENSIONS ALCP (LENGTH) AND BCP (BREADTH) THE ORIGIN OF XCP AND YCP IS AT THE INTERSECTION OF WALL1. A(I) = O.D0 BCP(J)/2.DO BCP(J)/2.PAREA C WRITE(6. SQFT = \F10.F12) FSCP(5.2) C DO 171 J = 11.'REMAINING CEILING AREA.6 FSCP(I1.D0 = HT CALL SHPRL (A.NCP+10 FOR PANEL NUMBER J.DO ALCP(J)/2.D0 B(2) = XCP(J) + ALCP(J)/2.B.ALCP(J)/2.J) DO 166 1 = 1 .25 DO 205 12= 1.NCP+10 178 PAREA = PAREA + ALCP(K)"* BCP(K) C NOW PRAREA IS THE AREA OF THE PORTION (OF THE CEILING ) WITHOUT PANELS PRAREA = (ALTH* BTH) .G. J) ARE CALCULATED C-32 .ALCP(J)/2.F12) FSCP(3.D0 C A(2) = BTH A(4) = HT C CALL SHPRP (A.D0 ) C A(2) = ALTH A(4) = HT C CALL SHPRP (A.DO C C AT THIS STAGE (ALL SURFACES.D0 C A(2) = ALTH A(4) = HT C CALL SHPRP (A.4 A(I) = O.BCP(J)/2.F12) FSCP(1.( YCP(J) + BCP(J)/2.J) = F12 C C FOR FSCP(4.B. A(I) = O.J) = F12 C C FOR FSCP(6.J) = F12 C C C FOR FSCP(2.DO C B(l) = XCP(J) .4 A(I) = 0.J) = O.DO 170 B(I) = O.D0 B(4) = XCP(J) + ALCP(J)/2.J) DO 170 I = 1.D0 B(2) = YCP(J) + BCP(J)/2.DO C B(l) = YCP(J) .D0 168 B(I) = O.4 .( YCP(J) .D0 B(3) = XCP(J) .ALCP(J)/2.F12) FSCP(2.BCP(J)/2.F12) FSCP(4.DO 169 B(I) = O.ALCP(J)/2.( XCP(J) + ALCP(J)/2.B.D0 B(2) = YCP(J) + BCP(J)/2.BCP(J)/2.D0 ) B(4) = BTH .B(4) = YCP(J) + BCP(J)/2.J) = F12 C C FOR FSCP(3.D0 B(2) = XCP(J) + ALCP(J)/2.B.B.J) DO 169 I = 1.D0 B(3) = BTH .( XCP(J) .DO C B(l) = YCP(J) .D0 B(3) = ALTH .D0) B(4) = ALTH .J) FSCP(6.D0) C A(2) = BTH A(4) = HT C CALL SHPRP (A.J) DO 168 I = 1. J.6 C WRITE(6.J).') = '.FSCP(J.413) C413 FORMAT (1X//.K=1.6) C-33 .192) (K .9X) ) C C DO 193 1=1.DO ENDIF C***** PRINTING OUT RESULTS C WRITE (6.I3.'PANEL-TO-WALL SHAPE FACTORS'.6(E10.K) = FSCP(K.FSCP(K.' .6) C194 FORMAT (1X/..TFSCP C ANGLE OR SHAPE FACTOR ALGEBRA .K1) = O.GT.6 C WRITE (6.P'. K=l.T15.NCP+10 180 TFSCP = TFSCP + FSCP(K1.9X) ) C C DO 212 1=1.J)* RAREA(K)/( ALCP(J)*BCP(J) ) CONTINUE C C 171 C C C CONTINUE WITH THIS ALL PANEL-SHAPE FACTORS ARE CALCULATED.6) C199 F0RMAT(1X/.'FSCP(P*.I3.(FS(I. J=11. IF REMAINING CEILING AREA IS GREATER THAN ZERO IF ( PRAREA.T35.E10.DO ) THEN C TOTAL OF FSCP(WALL OR FLOOR.6) .194)I.SUBROUTINES FSCP(6.. C &//. C & 'WITHOUT THE CEILING PANELS'.NCP+10 ) C412 F0RMAT(1X/.C C TO CALCULATE.I3.K=1.(T5. DO 181 Kl = 1.(FSCP(I.'WALL-TO-PANEL SHAPE FACTORS'.2X) ) C193 CONTINUE C C WRITE(6.5 TFSCP = O.6(I3.6(I3.6) C211 FORMAT (1X/.SURFACES) 179 DO 179 K = 1. (J.6) C192 FORMAT (1X/.K=1.') = \E10.I3.K.199) (J.PANELS IN CEILING) = TFSCP C TO CHANGE THE SHAPE FACTORS(SURFACES./) C DO 198 J = ll.NCP+10 C WRITE(6./) ) C198 CONTINUE C WRITE(6.5 FSCP(6.T35.4. O.5X.T5.T15.SEE ALGO. //.211) (K ./) C DO 201 K = 1.K=1.T5.T5.6 FSCP(J.'WALL-TO-WALL SHAPE FACTORS'.K) .REST OF CEILING).4.(T5.K).T15.195) C195 FORMAT (1X//.412) (K.FOR BHT.DO DO 180 LI = ll./) ) C201 CONTINUE C C TO COMPARE WITH THE EARLIER CASE C WRITE (6.6)* RAREA(K1)/PRAREA 181 CONTINUE C C ELSE DO 182 Kl = 1.6) = FSCP(K1.DO 182 FSCP(K1.6) = O.'FSCP('.I3.L1) FSCP(K1.* WALL-TO-WALL SHAPE FACTORS'.6 C WRITE (6.K1) = FSCP(K1.213)1.4.K) . F12) RETURN END C C C SUBROUTINE PARA(G.*)*B2= '.PQ(B2-A1.B.D2-D1) & .I3.G C WRITE(6.D1.D2-D1) RTOTAL = RHS1 + RHS2 + RHS3 + RHS4 C PI = DATAN(l.*)'C1= \C1 C WRITE(6.B(4) COMMON /COORDS/ Al.B2.T15.D2-D1) RHS4 = PQ(A2-B1.RTOTAL.*)'D2= ' D2 C WRITE(6. ' ATOTAL = '.*)'A2= '.6(E10.PQ(A2-A1.C2.C2-D1) .B(4) AND G (DISTANCE IN-BETWEEN ) C **** OUTPUT: F12 (SHAPE FACTOR FROM 1 TO 2) SUBROUTINE SHPRL (A.D1 C WRITE(6.*)'D1= '.C2-D1) + PQ(A2-B1.DO) * 4.D2 COMMON PDIST PDIST = G RHS1 = PQ(B2-B1.0-Z) COMMON /COORDS/ A1.PQ(B2-A1.Cl.DltA2.A2.Bl.B2 C WRITE(6.PQ(B2-B1.*)'B1= \B1 C WRITE(6.D2-C1) RHS2 = PQ(A2-A1.C2-C1) .PQ(B2-B1.D2-C1) & .C1.D2-C1) & .C2-C1) .F12) IMPLICIT REAL*8 (A-H.B1.C2-C1) + PQ(A2-A1.T5.C2-C1) + PQ(B2-B1.*)'C2= *.F12) IMPLICIT REAL*8 (A-H.B2.D2-D1) & .*)'A1= \A1 C WRITE(6.2X) ) CONTINUE RETURN END C C C C ** RADIATION SHAPE FACTOR F12 BETWEEN TWO PARALLEL SURFACES C **** INPUT : A(4).DO ATOTAL = 2.PQ(A2-B1.PQ(A2-B1.*) 'G = {.C2.D2-C1) RHS3 = PQ(B2-A1.4.C213 C212 C C C FORMAT (1X/.*)'RTOTAL = ' .A2 C WRITE(6.G.ATOTAL C-34 .C2-D1) + PQ(B2-A1.C2-D1) .DO * PI *(B1-A1)* (Dl-Cl) C WRITE(6.PQ(A2-A1.C2 C WRITE(6.O-Z) DIMENSION A(4).D2 C Al = A(l) Bl = A(2) CI = A(3) Dl = A(4) A2 = B(l) B2 = B(2) C2 = B(3) D2 = B(4) C G IS OBTAINED AS AN INPUT TO THE SUBROUTINE CALL PARA(G. C1.*)'PQ = '.PQ3 C PQ = PQ1 +PQ2 .F12) IMPLICIT REAL*8 (A-H.0-Z) COMMON PDIST C V = DSQRT( PDIST**2 + Zl**2 ) W = DSQRT( PDIST**2 + Z2**2 ) C PQ1= Z1*W* DATAN(Z1/W) C PQ2= Z2*V* DATANCZ2/V) C PQ3= (PDIST**2)/2 * DLOG( (W**2 +Z1**2)/(W**2) ) C C C C C WRITE(6. BETWEEN TWO PERPENDICULAR SURFACES INPUT:A(4).C1.C2.*)'PQ2 = '.PQ3 WRITE(6.B2.D1.C2.-C1) + RS(B2-B1.B(4) COMMON /COORDS/ Al.D2 C C Al Bl CI Dl A2 B2 C2 D2 = = = = = = = = A(l) A(2) A(3) A(4) B(l) B(2) B(3) B(4) C C CALL PERP(F12) RETURN END C C C SUBROUTINE PERP(F12) IMPLICIT REAL*8 (A-H.0-Z) DIMENSION A(4).PQ2 WRITE(6.D1.F12 = RTOTAL/ ATOTAL RETURN END C C FUNCTION PQ( Z1.*)'PQ3 = '.B1.A2.0-Z) COMMON /COORDS/ A1.*)'PQ1 = '.B2.D2.C2.A2.PQ1 WRITE(6.-D1) C-35 .B.B1.Z2) IMPLICIT REAL*8 (A-H.PQ RETURN END C C C C C C C RADIATION SHAPE FACTOR F12 .B(4) OUTPUT : F12 (SHAPE FACTOR FROM 1 TO 2 ) SUBROUTINE SHPRP(A.D2 RHS1 = RS(B2-B1. DO * PI *(B1-A1)* (Dl-Cl) WRITE(6.-C1) -i. IF (T .RS(A2-A1.-D1) RTOTAL = RHS1 + RHS2 + RHS3 + RHS4 C C C C C C C C C C PI = DATAN(l.RS(B2-A1.RS(B2-B1.*)'B1= '.C2.D2.C2.RS(A2-B1.*)'D1= '.*)'A1= ' .D2. * ATOTAL = *.-D1) + RS(B2-A1.0-Z) COMMON /HT/ G C T'= DSQRT( Y2**2 + Yl**2 ) C SPECIAL CASE: WHEN EITHER T OR Zl IS EQUAL TO ZERO.*)'B2= '.-D1) .D2.DO ATOTAL = 2.-C1) .RS(B2-B1.RS(A2-A1.D2.B1 WRITE(6.*)'A2= '.*)'PQ1 = '.D2.*)'C1= '.DO ) THEN RS2 = O.A1 WRITE(6.Y2.*)'PQ3 = '.Y1) IMPLICIT REAL*8 (A-H.25D0*( Zl**2 -T**2) * DLOG (T**2 + Zl**2 ) ENDIF C C WRITE(6.-C1) .EQ.C2.RS(B2-A1. O.*)'D2= '.ATOTAL F12 = RTOTAL/ ATOTAL END C C FUNCTION RS( Z1.DO .-C1) .C1 WRITE(6.OR.A2 WRITE(6.DO ELSE RS2= 0.C2.PQ3 C RS = RSI +RS2 C WRITE(6.D2.RS(A2-B1. O.D1 WRITE(6. O.EQ.-D1) RHS4 = RS(A2-B1.-C1) .EQ.EQ.*)'C2= '.D2 WRITE(6. IF (Zl . O.-C1) RHS2 = RS(A2-A1.*)'RS = \RS RETURN END C-36 .PQ2 C WRITE(6.C2.& & & & .*)'RTOTAL = *.RS(A2-A1. T .-D1) + RS(A2-B1. Zl .RTOTAL.C2.B2 WRITE(6.*)'PQ2 = '.DO ) THEN RSI = O.C2.AND.DO ELSE RS1= T*Z1* DATAN(Z1/T) ENDIF C SPECIAL CASE : WHEN BOTH Zl AND T ARE EQUAL TO ZERO.DO) * 4.-C1) RHS3 = RS(B2-A1.C2 WRITE(6.D2.DO .PQ1 C WRITE(6.-D1) .-D1) . APPENDIX D REPRODUCTION OF CHAPTER 8 FROM 1984 ASHRAE SYSTEMS HANDBOOK [ D-l tea . 8. Radiant panel systems are similar to other air-water systems in the arrangement of the system components (see Fig. suspended. walls or growth of the perforated metal. System Applications. reheat. Extruded panels can be manufactured in almost any shape and size. Lightweight metal panel ceiling systems respond quickly to load changes and can be used for cooling and heating. air or electric resistance. These radiant ceiling systems are usually designed into buildings where the features of the suspended acoustical ceiling can be combined with panel heating and cooling. The third type is an aluminum extrusion face sheet with a copper tube pressed into an oval channel on the back of the face sheet. Cold outdoor and heating medium temperatures must be analyzed with regard to potential damage to the building construction. or the panels can be arranged as large continuous areas for maximum economy. The air leaves the cavity through a normal diffuser arrangement and is supplied to the room. one-zone. The warm air system has a special cavity construction where air is supplied to a cavity behind or under the panel surface. its smaller mass enables it to respond more quickly to load changes. such as in an overhang. High temperature surface radiant panels [over about 250 F (121°C)] energized by gas.(900 x 1525 mm) and are held in position by various types of ceiling suspension systems. in buildings where glass areas are large and load changes occur faster. Radiant Space Healing & Cooling. in metal ceiling systems. these systems are used as floor radiant panels in schools and in floors subject to extreme cold. 1 Primary/Secondary Water Distribution System with Mixing Control The preparation of this chapter is assigned to TC 9. Panel Heating and Cooling Systems. One consists of lightweight aluminum panels. electricity or high temperature water are discussed in Chapter 18. the piping is accessible for servicing. Panel Cooling System Design ADIANT panel systems combine controlled temperature R room surfaces with central station air conditioning. Generally. The room heating and cooling loads are calculated in the conventional manner. Warm air and electric heating elements are two design concepts used in systems influenced by local factors. or it can include some or all of the features of dualduct. Modular panels are available in sizes up to about 36 x 60 in. usually 12 x 24 in. The controlled temperature surfaces may be in the floor. Room thermal conditions are maintained primarily by direct transfer of radiant energy. lag and override effect of masonry panels are unsatisfactory. The central station air system can be a basic. (12. acoustical ceiling. System Concepts. Because it is not likely to be covered. The panels can be designed as small units to fit the building module and provide extensive flexibility for zoning and control.5-in.CHAPTER 8 PANEL HEATING AND COOLING SYSTEMS System Types. However. and. constant temperature.88 x 1.7-mm) galvanized pipe coils. Three types of metal ceiling systems are available. SYSTEM TYPES Residential heating applications usually consist of pipe coils embedded in masonry floors or plaster ceilings.1 D-2 . attached in thefieldto 0. Electric heating elements embedded in the floor or ceiling construction and unitized electric ceiling panels are used in various applications for local spot heating as well as for providing full heating requirements for the space. The second consists of a copper coil metallurgically bonded to the aluminum face sheet to form a modular panel. Heat Transfer by Panel Surfaces. Panel Heating System Design. The ceiling panel systems commonly used today are an out- Fig. Also. A controlled temperature surface is referred to as a radiant panel if 50% or more of the heat transfer is by radiation to other surfaces seen by the panel. multizone or variable volume systems. rather than by convection heating and cooling. as are the floors. Practical limitations dictate a maximum size of about 16x4 ft (4. the slow response. This construe^ tion is suitable where loads are stable and solar effects are minimized by building design. and the temperature is maintained by circulating water.22 m). 1). Manufacturers' ratings gen- ceiling.1. Radiant panels are usually located in the ceiling because it is exposed to all other surfaces and objects in the room. This chapter is concerned with surfaces whose temperatures are controlled and are the'primary source of heating and cooling within the conditioned space. General Design Considerations. higher surface temperatures can be used. Radiant cooling can be incorporated. constant volume system. (300 x 600 mm). Thermal comfort. 5. it is increasingly difficult for convective systems to counteract the discomfort resulting from cold or hot walls. In much the same way that light energy from a lighting fixture illuminates the room so that all surfaces can be seen. AH pumps. The invigorating effect of radiant heat is experienced when the body is exposed to the sun's rays on a cool but sunny day in spring. Operable sash should be designed to discourage unauthorized operation. Maintaining the correct comfort conditions by low temperature radiation is possible for even the most severe weather conditions. Supply air quantities usually do not exceed those required for ventilation and dehumidification. floor or ceiling surfaces. Recessed lighting fixtures. This is significant because all surfaces within the room tend to assume an equilibrium temperature resulting in an even thermal comfort condition within the space. the air temperature does not vary at all. Comfort levels are better than those of other conditioning systems-because radiant loads are treated directly and air motion in the space is at normal ventilation levels. Heat transfer between the radiant panel and the other room surfaces is well established in a boxlike room where the primary heat gains and losses are from the wall. The mean radiant temperature (MRT) strongly influences the feeling of comfort. Draperies and curtains can be installed at the outside wall without interfering with the heating and cooling system. a rough surface emits heat rays more efficiently than a polished surface. 8. and an evaluation of this interrelationship is desirable. This becomes more critical if space dry." A person is not aware that his environment is being heated or cooled. Early evaluation is necessary to use the panel system to full advantage in optimizing the physical building design. Mechanical equipment is not needed at the outside walls. Heating and cooling panels neutralize these deficiencies and minimize excessive radiation losses from the body. 2. Cooling and healing can be simultaneous. should only be used in consultation with manufacturers experienced in this field. The maximum water treatment must not fuse the heads. The performance ratings presented in this chapter for radiation and convection can be applied directly to the calculated room heating and cooling loads. A common central air system can serve both the interior and perimeter zones. 4. where maximum cleanliness is essential or where dictated by legal requirements. Chapter 5. particularly those with large amounts of glass. a warm radiant panel emits energy that is absorbed and reradiated. each giving off radiant energy. Other factors to consider when using panel systems are: 1. Polished steel or similar polished surfaces show no such facets. Sections 5-6). (See NFPA 13-1982. which has been developed as a result of field testing. simplifying the wall. Other rays coming from the sun impinge on surrounding objects. [Inside single-glass surface temperatures . fans. panels should not be used in or adjacent to high humidity areas. Fortunately. the surface of concrete or rough plaster is covered with numerous facets. filters and so forth are centrally located. high velocity or high pressure drop devices or from pump and pipe vibrations.2 CHAPTER 8 erally are for total performance and can be applied directly to the calculated room load.8. This empirical information. 5. It is sometimes thought that a radiant heat system is desirable only for certain buildings and only in some climates.and wet-bulb temperatures are allowed to drift as an energy conservation measure. 12. although in such a short interval. wherever people live. 6. 4. floor and structural systems. and the inside surface temperature of glass is increased significantly. Every facet of the surface emits rays in straight lines at right angles to the facet. 9. 7. without central zoning or seasonal changeover. the performance of the radiant panel is related directly to the structure in which it is located. Various investigators and manufacturers report increased cooling performance because of solar effects and ceiling-mounted lighting fixtures. Some of these rays impinging on the body come directly from the sun and include the whole range of ether D-3 1984 Systems Handbook waves. reducing the potential for septic contamination. It is as important to provide the correct conditions in very cold climates as it is in moderate climates. 2. 13. Cooling. When the surface temperature of the outside walls. as defined by ASHRAE Standard 55-198/. As with any hydronic system. instantly there is a sensation of cold. In searching for the correct conditions compatible with the physiological demands of the human body. convection and evaporation. Thermal expansion of the ceiling and other devices in or adjacent to the ceiling should be anticipated. 10. most building surfaces have high emissivity factors and therefore absorb and reradiate energy from the active panels. The surface temperature of well constructed and properly insulated floors will be 2 to 3 deg F (1 to 2°Q above the ambient air temperature. 7. Research and testing of panel performance have been conducted by various independent researchers and manufacturers. where they are increased in wavelength and reflected to the body as low temperature radiation. these three factors of heat loss must be considered. when three. begins to deviate excessively from the ambient air temperature of the space. simplifying maintenance and operation. This feature is especially valuable when compared to other conditioning methods in existing buildings. and all surfaces become warm. The occupied space has no mechanical equipment requiring maintenance or repair. hung ceilings and other ceiling devices must be selected on the basis of providing the maximum ceiling area possible for use as radiant panels. The panel system can use the automatic sprinkler system piping. Panel heating and cooling systems function to provide a comfortable environment by controlling surface temperatures and minimizing excessive air motion within the space. The modular panel concept provides flexibility to meet changes in partitioning. When examined under a microscope. air diffusers. No space is required within the air-conditioned room for the mechanical equipment. producing a comfortable feeling of warmth. 3. SYSTEM CONCEPTS All bodies with a surface temperature above absolute zero emit rays with wavelengths depending on the body surface temperature. Should a cloud pass over the sun. 6.' is "that condition of mind which expresses satisfaction with the thermal environment. Thus. A 100% outdoor air system may be installed with less severe penalties in terms of refrigeration load because of reduced air quantities. hospital patient rooms and other applications where space is at a premium. The air-side design mast be able to maintain humidity levels at or below design conditions at all times to eliminate any possibility of condensation on the panels.and four-pipe systems are used. the piping system should be designed to avoid noises from entrained air. Unlike most heat transfer equipment where performance can be measured in specific terms. Warm ceiling panels are effective for winter heating because they warm the floor and glass surfaces by direct transfer of radiant energy. Wet surface cooling coils are eliminated from the occupied space. 11. However. Recent study has given us better insight on the human body and its response to the surrounding environment. Principal advantages of panel systems are: 1. no system can be rated as completely satisfactory unless it satisfies the three main factors controlling heat loss from the human body: radiation. 3. Panel Heating and Cooling Systems 8. thermally stable environment. and minimum air quantities delivered to the room are those required for ventilation and exhaust of the toilet room and soiled linen closet. peratures are restricted so as not to cause foot discomfort. The air supply system is often a 100% outdoor air system. The electric panels are easy to install and have the advantage of simplified individual room control. Other Building Types Metal panel ceiling systems can be operated as heating systems at elevated water temperatures. (2) requires no mechanical equipment or bacteria and virus collectors in the space requiring maintenance and (3) does not take up space within the room. With the increasing demand for worker comfort. where the walls and ceilings are cooled with chilled water. Often.or four-pipe concept may be used. there has been an increased application of electric resistance elements embedded in the floor or behind a skimcoat of plaster at the ceiling. dehumidification and usually some sensible cooling. Heating and cooling panel applications are similar to office buildings. Another advantage of panel heating and cooling for classroom areas is that mechanical equipment noise does not interfere with instructional activities. there has been only a limited application of this type in the Western Hemisphere. panel systems should be considered. In some applications. or for heating only. lobbies. If the school is air conditioned by a central air system and has perimeter heating panels. A single-zone central air supply system provides ventilation air. 7 are commonly observed. If cubicle tracks are applied to the ceiling surface. Cooling may also be applied. Besides absorbing heat from the space. Swimming Pools Panel heating systems are well suited to swimming pools because the partially clothed body emerging from the water is very sensitive to the thermal environment. Panel ceilings are often used in areas of the hospital occupied by mentally disturbed patients since no equipment is accessible to the occupant for destruction or selfinflicted injury. Water control valves should be in the corridor outside the patient room so that they can be' adjusted or serviced without entering the room. SYSTEM APPLICATIONS Apartment Buildings For heating. All piping connections above the ceiling should be soldered or welded and thoroughly tested. not directly over the water. ceilings can be installed at any height and remain effective. one special application is an internal combustion engine test cell. The high lighting levels in television studios make them well suited to panel systems. the occupants work in relative comfort when 55 F (13°C) water is circulated through the ceiling and wall panels. Ceiling panels are generally located around the perimeter of. the panel ceiling also improves the acoustical properties of the studio. electric resistance panels and forced warm air panel systems have all been used. The panel-system is usually sized to offset the transmission loads plus any reheating of the air required. Panel systems are readily adaptable to accommodate most changes in partitioning. The panels are installed for cooling only and are placed above the lighting system to absorb the radiation and convection heat from the lights and normal heat gains from the space. Prefabricated electric panels have also proved advantageous. the pool. For heating only applications. Industrial Applications Panel systems have found wide application in general space heating for industrial buildings in Europe. The coils must be carefully positioned so as not to overheat one apartment when maintaining desired temperatures in another. pipe coils are embedded in the masonry slab. tempered air is supplied at a constant volume.] As a result. The highest ceiling installed for a comfort application is SO ft (IS m) above the floor with a panel surface temperature of approximately 285 F (141°C) for heating. Lightweight metal panel ceiling systems have been applied to residences. The ceiling construction is made more rugged by increasing the gauge of the ceiling panels and using . Room control is accomplished by modulating the water flow through the panel. A separate minimum volume dehumidified air system provides the necessary dehumidification and ventilation for each apartment. particularly in add-on rooms. However. in all areas except gymnasiums and auditoriums. Although the ambient air temperature in the space ranges up to 95 F (35°C).3 10 to IS deg F (5 to 8°C) above those indicated in Fig. Panel surface temperatures are higher to compensate for the increased ceiling height and to produce a greater radiant effect on the partially clothed body. the panels are arranged for zone control. This system is well suited because it: (1) provides a draft-free. Hospitals The principal application of radiant panel systems over the past 30 years has been for hospital patient rooms. The ceiling panels offset the heat loss from a single-glazed all-glass wall. using grid coils in the floor slab or copper tubing systems in older plaster ceilings. The embedded pipe coil system is the most common. a single-zone piping system might be used to control panel heating output. and the room thermostat would modulate the supply temperature or supply volume of air delivered to the room. Office Buildings The panel system is usually applied as a perimeter system providing heating and cooling.or four-pipe design. Individual room control is usually by throttling the water flow through the panel. The slow response of embedded pipe coils in buildings with large glass areas may prove to be unsatisfactory. Electric panels in lay-in ceilings have been used for full perimeter heating. The piping system may have a two. and have been used in airport terminals. convention halls. Schools Panels are usually selected for heating and cooling. Because radiant energy travels through the air without wanning it. Installations for heating and cooling have been made with pipes embedded in a hung plaster ceiling. the system may be used with any type of approved ventilation system. In recent years. museums and especially where large glass areas are involved. and the room thermostat modulates the panel output. Installation with ceiling heights of SO ft (IS m) and single glass from floor to ceiling provide satisfactory results. Floor panel tem- D-4 Residences Embedded pipe coil systems. Water distribution systems using the two. track installation should be coordinated with the radiant ceiling. and the air system is designed to provide individual room control. Installations can be made where complete flexibility is on a modular basis. Metal panel ceiling systems are also installed in minimum and medium security jail cells and other areas where disturbed occupants are housed. downdrafts are minimized to the point where no discomfort is felt. These systems are well suited to normally constructed residences with normal glass areas. For example. In practice. the combined factor is about 0. there are many factors that interfere with or affect natural convection. F (°C). (2). which is a variation of Eq. (2)] is used frequently. that the equation is accurate to within 10% when used in conventional heating and cooling calculations. In the following paragraphs. such as shipboard berthing spaces with an adjacent hot gas stack. BTUHPERSO'T 120 ISO WO ISO Fig. In confined situations or special applications. Radiation removed by a cooling panel for a range of normally encountered temperatures and as calculated from Eq. Tests4 show that the value of the constant of Eq. and the equation for heating can be rewritten: or for cooling: D-5 0 O 20 JO 40 SO 60 TO 60 90 100 MO HEM OUTPUT. the room temperature at the 5-ft (1. This equation assumes a simple. and all surfaces are perfectly diffusing. The values apply to ceiling. the two transfer mechanisms are first considered separately and then combined to facilitate design calculations. A i and .0. Convection Transfer The convection coefficient qe is defined as the heat transferred by convection in Btu/h«ft2-F (W/m2-°C) difference between air and panel temperatures. The configuration of the room and the spaces determines the natural convection. Floor or Wall Panel where qr =heat transferred by2the panel to or from the room surfaces by radiation.1713 = Stefan-Boltzmann radiation constant. Infiltration.°Q absolute temperature to the fourth power.42 = areas of the surfaces. Fe = the emissivity factor (dimensionless). the wall surface temperatures tend to rise considerably above the room air temperature.1S. it is necessary to evaluate each surface using the geometrical factors from the charts in Chapter 2 of the 1981 FUNDAMENTALS VOLUME.1S2 in the test room. tp =the average panel surface temperature. Btu/h«ft2«F (5. Btu/h«ft (W/m2). Substituting this value in Eq. Tr = mean radiant temperature of unheated surface.8. Btu/h«ft2 (W/m2). 0=460(273). the movement of persons and mechanical ventilating systems can . It is generally agreed. these refinements are generally insignificant. F (°C) abs. (3a). F|. 3. or in situations where the emissivity of the surfaces is significantly different. when considering spot cooling it may be necessary to consider the influence of gaseous radiation. the emissivity of nonmetallic or painted metal nonreflecting surfaces is about 0. When this emissivity is used in Eq. however. The Hohel equation [Eq. Where several surfaces exposed to the panel have widely differing temperatures. The design information in this chapter is based on that constant value. (3) is given in Fig. all other surfaces are at another temperature. Radiation exchange calculated from Eq. floor or wall panel output. Tp = average surface temperature of heated panel. 0. 2 Heat Transferred by Radiation from a Heated Ceiling. (1). The actual radiation transfer in a room may be somewhat different from that given by Eq. that is. In structures where the main heat gain is through the walls or where incandescent lighting is used. (3). irregular room surfaces. Room related angle and shape factors can also be found in Fanger's book on thermal comfort2 or from the algorithms in ASHRAE Energy Calculations 1.9. F(°C). AUST =area-weighted average temperature of the unheated surfaces in the room. 2 + [(l/e I )-l)+yi l />l 2 [(l/e 2 )-l] where Fc = combined configuration and emissivity factor.87 for most rooms. F =F F = (2) c ' r ' l/F 1 . 2.2 == view factor = 1. the constant becomes about 0.3 For normal application. This equation may also be written as: q. floor or wall. Fa = the configuration factor (dimensionless).4 CHAPTER 8 1984 Systems Handbook security dips so that the ceiling panels cannot be removed. (3) or (3a) because of nonuniform temperatures. is given in Fig. In practice. boxlike room in which there is a uniformly heated ceiling.1976. however. F (°C) abs. HEAT TRANSFER BY PANEL SURFACES A heated or cooled panel transfers heat to or from a room by convection and radiation. =0.ni3FaFe [(7V/100)4 -{T„/\00)A] (1) where qr = heat transferred by radiation. et ande2 emissivities of the surfaces. In many specific instances where normal multistory commercial construction and fluorescent lighting are used. variations in emissivity of materials and so forth. it may be necessary to-compute the area-weighted Average Unheated (or Uncooled) Surface Temperature (AUST) exposed to the panels. Heat transfer convection values are not easily established. Part of the perforated metal ceiling can be used for air distribution. Radiation Transfer The basic equation for radiation exchange is the StefanBoltzmann equation (see Chapter 2 of the 1981 FUNDAMENTALS VOLUME).6697 x in -8 W/m2 . Convection in panel systems is usually considered to be natural. Similarly. air motion is generated by the warming or cooling of the boundary layer of air which starts moving as soon as its temperature rises above or drops below the surrounding air temperature. (3) and (3a) was 0.5-m) level will closely approach the AUST (Average Uncooled Surface Temperatures). . 3 Heat Removed by Radiation to a Cooled Ceiling or Wall Panel introduce some forced convection that will disturb the natural process. tp = temperature of panel surface... The most consistent results are obtained when the air layer temperature is measured close to the region where the fully developed stream begins.. The convection in a panel system is a function of the panel surface temperature and the temperature of the airstream layer directly below the panel. 4 Heat Output by Natural Convection from Floor and Ceiling Heating Panels Natural convection from a heated ceiling qc = 0.._.26 (t„-tay» (9) Figure 4 shows heat output by natural convection from floor and ceiling heating panels as calculated from Eq. Figure 5 shows heat removed by natural convection by TEMPERATURE DIFFERENCE TO PANEL SURFACE-F DES Fig. / > / / / /1 .5 m) above the floor in a 12 x 24 ft (300 x 600 mm) room. OF AIR (Ip .7 Other tests8 established that the effect of room size was also usually insignificant.*.6 Research4 has determined natural convection coefficients referred to the center of the space 5 ft (1.ta)125 (7) Natural convection from a heated floor or cooled ceiling qe = 0 . Equations (4) to (9). F iy / s 90 / _. — ! « —— — — — XV..' * . . ta = temperature of the air.5 in. ft(m)..*. I 0 SUM O F At L SURFACE S-090 SO / / 1 ..)'•"/£>. . derived from this research.Panel Heating and Cooling Systems 8. because the increments are unpredictable in pattern and performance and do not significantly increase the total capacity of the panel system.39(/ p -/ f l ) "/£>... ft (m). (SI to 64 mm) below the panels. a 3 1- fc X' 1 f / in . usually 2 to 2.tayM/H00i (6) "where qe = heat transfer by natural convection. The convection equations can therefore be simplified to: /.. Measurements of panel performance in furnished test rooms that did not have uniform temperature surfaces showed variations that are not large enough to be significant in heating practice..5 eo 1 i 1 7S 1 1 CONVf RSION FACTORS: 6S • j i 1 COMVIRSIOKI FACIIM3.F/i. The effect of forced convection on heat transfer from panels has been reported3 as an increment to be added to the natural convection coefficient. F (°Q. can be used to calculate heat transfer from panels by natural convection. X IS 60 70 PANEL SURFACE TEMPERATURE . However.29{tp ."•* (4) Natural convection from a heated floor or cooled ceiling .. — / .041 (f„-/./ J .021 (tp .. c..-. although some of the pioneer work has been done. : " (8) Natural convection from a heated or cooled wall panel qr = 0.r. OF PANEL SURFACE MINUS TEMP. — .. • S 10 IS 20 25 30 31 401 4 5i SO 1 SS 60 65 TEMP. ^ --. (7) and (8). .a Wta 1 > BWft I r X l B ~ 1 1 s50 =III 4 » - N>. D-6 . - b" \«/._~ .. Natural convection from a heated or cooled wall panel qc = 0. increased heat transfer from forced convection should not be used. Very little heat transfer literature describes experiences pertinent to this application. H «* height of wall panel.. Natural convection from a heated ceiling qc = 0.lo) - 70 F Fig.: 008 (5) 9 c =0. Btu/h» ft2 (W/m2). 3 2 ( / p .. • i * a 60 V / f* V'EW FAC ' O R . F (°C).* & • rTLT — iao»"o»j criiMiB WMtia-V 0 Fig. 5 Heat Removed by Natural Convection to Ceiling Cooling Panels . De = equivalent diameter of panel (area X4T perimeter). radiation from the top of the fixtures will raise the overhead slab temperature and will transfer heat to the ceiling space by convection. 1. 7 Inside Wall Surface Temperature Correction for Air Temperatures Other Than 70 F Effect of Floor Coverings Floor coverings can have a pronounced effect on the performance of a floor heating panel system. 10. depending on the type of bond between the water tube and the panel material. for perforated metal panels.2 Fig. corrosion between lightly touching surfaces. it may be possible to maintain a high enough water temperature to satisfy the covered panels and balance the system by throttling the flow to the bare slabs. In this case. O 0. however. heat can be transferred from the ceiling panel to the floor slab above (heating) and vice versa (cooling). the increased water temperature required when car- D-7 . 3. as in Fig. to a much smaller extent. Refer to the manufacturer's data to include these heat gains. 6 for a 70 F (21°C) room air temperature. 5. which. lights. the temperature of the ceiling space should be determined by testing. 3 and 5.6 40 SO 20 OUTDOOR AIR IO O -10 TCMPCRATURC -OCCRCC FAHR. As a preliminary basis for design in Fig. also provides acoustical control.B T U PER ( H « ) ( S Q F f X f DEC) 1. Use of Fig. If lighting fixtures are recessed into the suspended ceiling space. In calculating the AUST. (4). is shown in Fig. An additional curve is shown to illustrate the effect of forced convection on the latter data.68 W/m2«°C) of temperature difference between room design and panel temperature. This bond may change with time. Corrections for other temperatures may be obtained from Rg. method of maintaining contact and other factors.7. since it varies with different types of panel systems. The radiation component can be approximated using Fig.some instances. In suspended ceiling panel systems. 2 or 3. In. as given in Fig. Data on the thermal resistance of common floor coverings are given in Table 5. -20 [Indoor air Hmptraturo —70 f\ Fig.0 IJ O V E K A U . again depending on configuration and insulation. the surface temperature of 1984 Systems Handbook the inside walls is assumed to be the same as the room air temperature. The surface temperatures of outside walls and exposed floors or ceilings for heating panel calculations can be obtained from Fig.2 0. people or equipment. but the effects would be of the same order of magnitude.10 show that the temperatures are almost equal. 8 and 9. 10 do not include heat gains from sun.7 OB Ofl 1. Tests9. Theoretically. use 1 Btu/h«ft2«F (5. Similar adjustment of the ASHRAE data is not exactly appropriate. Most system manufacturers have empirical information available. This energy will be absorbed at the top of the cooled ceiling panels by radiation. The actual thermal resistance of any proposed system should be verified by testing whenever practicable. generally in accordance with Eq. 2 or 3 to the convective heat transfer from Fig. The combined heat transfer for ceiling and floor panels when used for heating in rooms in which the air temperature is 70 to 76 F (21 to 24°Q can be read directly from Fig. 4 or 5. 2 and 3 requires calculating the AUST in the room. The combined radiation and convection transfer for cooling. the temperature difference used is that between the top of the ceiling panel and the midspace of the ceiling. the temperature of the heating medium must be increased. The amount the top of the panel absorbs depends on the system type.CHAPTER 8 8. respectively.5 0. and by convection. To maintain a given upward heat flow after a floor covering has been added. The data in Fig. Panel Thermal Resistance The thermal resistance to heat flow may vary considerably among panel systems. (8) and data6 for specific panel sizes. The ceiling panel surface temperature is affected because of heat transfer to or from the panel and the slab by radiation and.1 0. These two diagrams apply to rooms in which the AUST does not differ greatly from room air temperatures. However.6 0. COEFFICIENT OF HEAT TRANSFER U . panels'installed under a roof will absorb additional heat. The added thermal resistance of the floor covering reduces upward heat flow and increases the heat flow to the underside of the slab. respectively. Tables 1 through 4 show some typical values for thermal resistance factors for various types of floor and ceiling panels. Similarly. Where covered and bare floor panels exist in the same system. Combined Heat Transfer (Radiation and Convection) The combined heat transfer from a panel to a room can be determined by adding the radiant heat transfer from Fig. much of this heat transfer is nullified with the application of insulation over the ceiling panel. 6 Relation of Inside Surface Temperature to Overall Coefficient of Heat Transfer cooled ceiling panels as calculated by Eq. by convection.3 O * 0. The convection component can be approximated using Fig. Panel Heating and Cooling Systems 8.3 W . ^-^ ^~ -—-^1. Output Upward peting is applied over floor panels makes it impossible to balancefloorpanel systems in which only some rooms have car- D-8 peting. 9 Floor Panel Design Graph Showing Pane) Surface Temperature and Mean Water Temperature vs. I * * ^j i 1 OUIH .J ^ ro r^ • I3> jf- i " */ CONV ERStON FACTORS *C * ( F . / ' .? ^ • ^ iS ^ I ^ ^ 8 Cm I / ^ €%-^ ^ ^ -^ P^ ^ •^s* ***^./ y y X S00^ . M.M..7 X 'X . 8 Ceiling Panel Design Graph Showing Panel Surface Temperature and Mean Water Temperature vs. ^ c >sf^ 1.-C iW / / e tt'-F-tVBtu * (t value) ttiffi / / / / _*zL*iUA L ^ LftdL*JL«4LsA .#/ -*A-?A-#A-$A-&L$A-&L?A-*A. unless the pipe is arranged to permit zoning using more than one water temperature. .B o/. *l L^.^Z / / / SURFACE OR MEAN WATERTEMPERATURE — F Fig. Output Downward ^^VV^VVVVv^^^VV^VVsVV^^^Jî^' SURFACE OR MEAN WATER TEMPERATURE — F Fig. 09) I-in.SI 2.25 (228. check floor surface temperature for possible foot discomfort (see Ref 8).16) (0.73 0.1 (0.07) 0. (12.08) a Any ceiling panel also acts as afloorpanel to the extent of its upward heat flow.08) (0.I 0.4 m a ) Cover i 70 Thermal Resistance.19) (0.) non-ferrous tube 9 (228.92) (0.75-in.51 2.8) 0.11) (0.08) (0.79) 0.9 0.56) 0.6) 12 (304.°C/W) Heal Flow Hallo.04) 0.45 (0.5 (0.63) 0.20 0.40 0.09) (0.44 0.6 (0.06) (0.19) (0.12) (0.12) (0.3 (0.30 (0.8 CHAPTER 8 1 Table 2 Thermal Resistance of Concrete Ceiling Panels (Heating) 1 I CONVERSION FACTORS: •ft'xais W/m'-BtuA .07) (0.08) 9 (228.93 0.21 0.32) 0.7 (0.07) (0.83 0.8) (0.20) (0.45 (0.25 0.11 (0.17 0.04) 0J0 0.43 2.37) 0. These coefficients should not be used to determine the downward heat loss from panels-built on grade be- .14) (0.84 0.45 2.8) 1.58) 4J (0.12) 15 3.16 0.75-in.48 1.07) (0./qa.4 mm) Concrete Slab— M a .29) (0.6) 12 (304.11) (0.50 0.6 (0.63 0.36 (0.16) (0.08) 0.l mm)(nom.6) 12 (304.F Panel Coustracuon Spacing.04) 0.6 mm) Concrete Slab— 24n.05) 0.18) (0.10) (0.6 0.90 0.46 0.18) 0.4 mm) (nom.1 0.05) (0.49 0. qu/arf org. (nom.08) (0.05) 0JS 0.8 (0.63 0.17) (0.50 1.35 (0.07) (0.) non-ferrous tube 10 down 3.9 0.09) (0.6) 12 (304. SA.7 0.08) (0.) nonferrous tube 9 3.49 2. In either case.35 (0.ll)(0.30 (0.04) (0.8 (0.23 0.(l9.34) 0.10) (0.20) (0. (12.5 0.16) 2.05) 0.23) (0.J0) (0.7 0.7 mm) (nom.09) (0.) ferrous pipe 9 (228.0 (0.09) (0. (15.6) 12 (304.08) (0.9 0.40 0.46 0.) ferrous pipe or0.22) (0.78 0.11) (0.42 0.05) 0.76) (0.8 0.12) (0. (12.61 0.35 (0.09) (0.57 0.05) 0.0 (0.7 mm) (nom.57 1. (203 J mm) Concrete Slab— Wo.28) 2.78 0.6 0.16) 0.07) (0. In.43 1.08) S-ln.59 0.30 0.16) (0.30 (0.19) (0.06) 12 4. «-ta.4 mm) (nom.12) (0.58 1.7S4n.05) 0.07) (0. If the heat is transferred to another heated space.13) (0.16) (0.7 0.54 3.) nonferrous tube 9 (228.75-in.07) (0.14) 1.(l2. Panel heat loss to space outside the room should be kept to a reasonable amount by insulation.754n(l9. (12.78 0.14) (0.05) 0.) nonferrous tubcorl-in.05) (0.10) (0.60 1.06) (0.18) 1.12) (0.05) 0J5 (0.14) (0.30 (0.25 (0. it is not good practice to have the major portion of the upper room's heating requirements supplied by the upward heatflowof a ceiling panel below.1 0. heat losses are part of the building heat loss if the heat is transferred outside of the building.80 (0.30 (0.9 0J5 (0.40 (0.09) (0.21) (0.lmm) (nom.21 1.25 (0.45 5.20) (0.9 (381) (0.90 0.0 0.97 (0. For example.73 0.7 0.) non-ferrous tube 9 (228.40 (0.(12.) ferrous pipe 12 (304.08) (0.06) 0.0 0.25 1.55 (304.12) (0.07) 12 5J 0.10) (0.61 (0.30) (0.54 1.46) 0.25 (0.08) (0.4 mm) (nom.15) (0.18) 0.51 0.07) (0.30 0.30 (0.35 1.) ferrous pipe 'us up 0.98 0.6) (0JI) (0.48) D-9 down np down 'us 'd 'us 'd 'us 'd 0.05) (0.12) (.08) (0.7mm)(nom.30 1.04) 0.39) 3.45 (0.8) 0.7 0.58) 4.63) (0.41) 6-tn.8) 15 (381) 0.6 (0.07) (0.68 0. (nom.10) 0.08) (0.5-in.7 0.8 mm) Oner 0.04) 0.45 1.09) (0.08) (0.) nonferrous tube 9 (228.50 (0.8) 2.66 0.04) (0.11) 0.9 (0.19) (0.40 (0.04) 0.08) (0.08) (0.07) 0.31) 0. ($0.30 (0.47 0.46 I JO 0.63) (0.10 (0.04) 12 (304.21) (0.04) 0.7 (0.04) 4.35 0.7mm) (nom.70 0.) non-ferrous tube Thermal Resistance. (25. 1 qulqd 0.20 (0.45 (0.7$-in.08) (0.25 (0.39 0.5-in.05) 0.24) (0.48 0.11) (0.15) (0.70) (0.12) (0.70) (0.39 0.46) 0.50) <L5-ln.38 1.. the underside of floor panels or the edges of any panel is considered a panel heat loss. f|2.04) 0.40 (228. the panel loss is a source of heat for the space and is not a part of the building heat loss.8 0.6) 12 (304.07) (0. (25.71 0.15) (0.14) (0.) ponferrous tube 9 (228.25 0.) ferrous pipe or 0.38 1.10) (0.09) 0.07) 0.7 0.35 ((0.87 0.14) (0.46 2. If the upward heatflowis high and the space above is occupied.41 1.3 0.06) (0.05 0.7 0.74 (0.(l9.25 (0.52 0.8 mm) Cover TO 03 up 9 (228.05) 0.20 (0. (152.) nonferrous tube 9 (228. the magnitude of panel loss should be determined.4 mm) Cover 0.06) 0.8. (25.10) 9 2.68 0.35 1.10) (0.8 0.8) (0.5-in.35 (0.16) (0.05) 0.8 (0. (19.10) (0.30 0.37) 3J (0.46) (0.06) 0.47 1.48) 3.56 (0.77 0.54 1.s _r l\ • V • THICKNESS PANEL .75-in.) ferrous pipe or 0.12) (0.40) 0.07 0.14) (0.66 0.16) (0.6) (0.10) (0.4 mm) Coacnrie Slab— 14a. No Infiltration.0 (0.04) 0.07) 0.6 0. (25.21) (0.30 (0.8) 0.07) 0.90) (0.42) 1-to.1 (0. Panel heat loss to heated spaces may require reduction by insulation if the amount of heat transferred is excessive or if objectionable temperatures develop.40 (0.12) (0.8 0.40 (0.8) 0.9 0. \ ^ COVER—" 75 Fig.14) (0.63 1.07) 0.2 0. Also check effect on heating requirements of the space above.09) (0.) ROnfcrroustube CEILING PANEL TEMPERATURE.9 0.7 mm) (nom.08) 0. the back surface of wall panels. hi.29) (0.6 0.8) 2. (SO.45 (304.50 (0.b/Bto (m*.2 (0. 10 Performance of a Cooled Ceiling Panel (Uniform Environment.10) (0.86 (0.7mm) (nom.8) 0.08) 15 44a.40 0.16) (0. Panel Table 1 Thermal Resistance of Bare Concrete Floor Panels (Heating)* Spacing.5-in(l2.09) (0.35) (0.30 (0.06) 0.6 0.06> 0.30 I.40 (0.8) IS (381) 2.97 0.73 0.51 1.45 (0. 11* • F • h/Btu (m2 .~ Kent Flow Ratio. a floor panel may overheat the basement below and a ceiling panel may cause the temperature of a floor surface above it to be too high for comfort unless it is properly insulated.11) (0.08) (0.10) (0.30 (0.6 0. (19.06) (0.7 ram) (nom. The heat loss from most panels can be calculated by using the coefficients given in Chapter 20 of the 1981 FUNDAMENTALS VOLUME.05) (0.6) 12 (304.12) (0.6) (0.25 (0. (152. (25.48 2.43 1.35 (0. No Internal Heat Sources) Panel Heat Losses Heat transferred from the upper surface of ceiling panels.0 0.16) 1.4 mm) (nom.6) 12 (304.39 0.07) (0. (101.26) 0.8) 15 (381) 0.5-ta.14) 0.10) 1-in.12) (0.59 0.16) (0.08) 0.54 1."C/W Panel Construction x\ 5 1984 Systems Handbook 9 1.40 0.1 mm) (nom.40 (0.1mm) (nom.) ferrous pipe or I-in.41 0.12) (0.06) 0.55 0.55 (0.SURFACE .1 mm) (nom.07) (0.20 (0.9 (228.4 mm) (nam.16) (0.6) 12 (304.55 0.18) (0.9 0J0 (381) (0.14) (0.) ferrous pipe or Inn.09) (0.9 (0.18) (0.10) (Q.08) (0.65 0.) ferrous pipe 9 (228.F.16) (0.7 (0.25) (0.6) 12 (304.14) 1.35 (0.0 0. (19.63) (0.23) (0. (25.8) 0. (nun) 1 3 5 up dowa up down np down up dowo 'ds 'us 'ds 'us 'ds 'us 'ds 0.13 0.98 0.13) (0.17) (0.1 mm) (nom.5-in.8 (0.42 1. 2 0.75 (0-13) 1. walls or floors.31" (0.12) 1. cause the heat flow from them is not uniform.9 III! STEEL PIPE 0. or as a cooling panel when chilled water is flowing.61" (0.) 6 ferrous pipe above metal lath tied at (192) 8-in. a source of dehumidified ventilation air is required in summer. and the system is classed as one of the combination air-water systems.4 0.57 (0. Figure 14 illustrates a metal panel ceiling system using copper tubing.11) 1.43 (0.) nonferrous tube or '/S-in.8 1. Also check effect on heating requirements of the space above. 4.140F(60°Q.0» 0. 12. ft 2 .* -r— T 1. Electrically heated ceilings or floors. acting as a heating panel when warm water is flowing.45 (0.05) qu/qj 0.4 1. 11 Downward and Edgewise Heat Loss Coefficient for Concrete Floor Slabs on Grade Metal Ceiling Panels Metal ceiling panels are usually integrated into a system that heats and cools.B t W I i r i * fXSJBB (C. Electric ceiling panels.0 i 1 METAL LATH ¥ ii METAL OR 1.30 (0.30) 0.17) 1.3 J—*-H I N S U L A T I O N AT SLAfc/ EDGE ONLY d .0 1. rd.06) 0.I / G Y P S U M LATH i •m Mmm>m^<&i 1.5 mm) (nom.5-in. 11.15 (0. depending on local codes.05 light carpet Light carpet with rubber pad Light carpet with light pad Light carpet with heavy pad 0.01) 0.15 (0. It is not good practice to have the major portion of the upper room's heating requirements supplied by the upward heal flow of a ceiling panel below.46 (0.09) 0. metallurgicaliy bonded to an aluminum panel that can be mounted into various types of ceiling 12 (305) -ALUMINUM PAN 0 .""13 The heat loss from panels built on grade can be estimated from Fig.8 1.or 600-mm) centers. bRecommendcd maximum inlet water temperature ('max) . (150-.» 0. (102) 8 (203) 0.2 1. (12.38 (0.08) 0. on either 6-.33) •Any ceiling panel also acts as a floor panel to the extent of its upward heat now.071" (0.29 (0.51 (0.6 0. Also.05) TUBE SPACING | i a Resistance fruef ft*«F«h/Bto <m2»°C/W) Bare concrete.15) 1.20) 0.25 (0.5 (114) TUBES OR PIPES Heat Flow Ratio.3 INSULATION CONDUCTANCE BTU PER (HR) ISO FT) (F DEC) 0.69 (0. Table 4 Thermal Resistance of Metal Ceiling Panels Thermal Resistance to Heat Flow fti'F'h/Btu (m»»0C/W) Type of Panel Spacing—Inches (mm) PAN E 0 6 E HELD AGAINST PIPE BY SPRING CUP 3 (76. 3.57 (0.03) •r 1 FACTOAfl XO3048 W/m 1 .08) 0. Metal ceiling panels.2 mm) intervals 9 with good tube embedment. Figure 12 illustrates a metal ceiling panel system that uses 0.23) 0.00 O.F>h/Bra (m 2 . (9.OS 0. Three Coats SPACING •:-. EITHER WAV Spacing.34 (0.6 1. (9.7 1 O.5 mm) (nom.28) 0.10) 0. no covering Asphalt tile Rubber tile 0.11) 4 COPPER TUBE SOLOEREO TO 0 .95 (0. In.7 (0.09) (mm) 4.71 (0. P A N E L HEATING A N D COOLING SYSTEMS The most common forms of panels applied in panel heating and cooling systems are: 1.07) 0.42 (0.6 1. check floor surface tempciature for possible foot discomfort (see Ref 10). ^TUBES < o \ 0.07) 0. 0 4 0 IN.25) (0. THICK ALUMINUM SHEET •Manufacturer's data. Plaster Pancb b Standard 0.or 24-in.C .50 (0. Aluminum ceiling panels are clipped to these pipe laterals.0 ^CoOj. 0 3 2 IN. THICK \ comfERStON Table 5 Thermal Resistance of Floor Coverings Description 0.05 (0. 2.) nonferrousl6 tube below m clal or gypsum lath (406) .2 1 (NO INSULATION C 3 =I.63 (0.7 mm) (nom. (229) or 3/8-in. a ventilation system may or may not be required.25) 0.B) I.11) (0.2) 6 (132) 0. 5. hydraulically connected in a sinuous or parallel flow welded system.4 Fig. Air-heated floors.20) 1. Embedded piping in ceilings.18) 1.22) 1. 300.15" (0."C/W) Gypsum Plaster.75 (0. (13-mm) pipe laterals.I 1 0. (203.85 (0.2 0.8 D-10 !!! 3/8-in.85 (0. When metal panels are applied for heating purposes only.9 4 0. If the upward heat flow Is high and the space above is occupied.9 ySLAB rOUNDATK Table 3 Thermal Resistance of Piaster Celling Panels (Heating or Cooling) GRADE Thermal Resistance to Downward Heat Flow. various amounts of forced air are supplied year-round.18) (0. In such a system.13) 1.30) Heavy carpet Heavy carpet with rubber pad Heavy carpet with light pad Heavy carpet with heavy pad 0.Panel Heating and Cooling Systems 8. 8°C) rise for cooling.10 CHAPTER 8 . 12 and 13 is designed so that the grid system can expand relative to the suspension system. Walls and Floors When piping is embedded in ceilings. The suspension system holding the ceiling grids in place must be braced by cross-furring against the walls of the building in both horizontal directions. while compatible matching acoustical panels are selected for the remaining ceiling area. If the lath is suspended to form a hung ceiling. the coils.-. 3. either on the back side of the panels or on the underside of the overhead floor system.„. generally within an inch (25 mm) of its lower surface. per 100 ft per 100 deg F (0. 16. but in good contact with.oy i -.: -.8 in. If plaster is to be applied to the concrete. 15 Coils in Structural Concrete Slab WIRE TIE T O SUPPORTS SCRATCH COAT HEATING / EMBEDDING PIPING / PIPES 3 COAT PLASTER FINISH SUPPORTING MEMBERS ON 3 OR 4 FT CENTERS. It is common to design for a 20 deg F (11°C) temperature drop for heating across a given grid and a 5 deg F (2. PIPE LATERAL PANEL CLIP Fig.o : ' -"°.o. Extruded aluminum-type panels are often used as long-narrow panels at the outside wall and are independent of the ceiling system. 16 Coils in Piaster Above Lath D-ll . .v-. the piping should be installed not less than 0.i. ' o ' . The acoustical blanket is also required for thermal reasons. MODULAR RADIANT AC0UST4CAL PANEL 3 COAT PLASTER FINISH I P L A S T E R C E I L I N G BELOW Fig.. Other forms of ceiling construction are composition board.75 in. but the suspension system should not move in relation to the building. 2. so that the reverse loss or upward flow of heat from the metal ceiling panels is minimized. The metal panel system shown in Fig. one of the following constructions is generally used: 1. ALUMINUM SHEET. 12 Metal Ceiling Panels Attached lo Pipe Laterals suspension systems. Smaller diameter copper tube is attached to the underside of wire lath or gypsum lath. The maximum design value of movement is 0. 17. H E A T I N G C O I L S " •.8. .-. The minimum coverage must comply with local building code requirements.• o : 'CONCRETE SLAB. Metal ceiling panels can be perforated so that the ceiling becomes sound absorbent when acoustical material is installed.and four-pipe distribution systems. 14 Extruded Aluminum Panel with Integral Copper Tube COftVf NSKMI ACTOR V . '" . Pipe or tube is embedded in the lower portion of a concrete slab.<>- o . o.'. Metal ceiling panels can be used with two. the piping may be placed directly on the wood forms.67 mm/m per 56°C). USE PIPE OR STEEL STRUCTURAL MEMBERS N J M E T A L LATH FINISHED PLASTER C E I L I N G SUSPENDED PLASTER CEILING SCRATCH COAT EMBEDDING P I P E S COPPER TUBE.". o. Such large grids should have metal panel ceiling expansion joints that line up with the expansion joints of the building..56 in. Figure IS illustrates a metal panel using a copper tube pressed into an aluminum extrusion. Figure IS shows this method of construction. '»:o:o\?: i \%-. 0 -'. as shown in Fig. Some ceiling installations require active grids to cover only a portion of the room. The expansion and contraction of the ceiling grids are compensated for by allowing the metal panels to move or adjust in the wall molding. Many large grids have been constructed in the field by butt welding the pipe laterals together to produce 100 ft. carefully embedding the coil as shown in Fig.v Fig. THERMAL «L*N*CT 1984 Systems Handbook SUPFOIIT BRACKET COPPER TUBE PRESSED INTO AND CUPS OVAL CHANNEL . A steel pipe grid will expand 0.?io. Embedded Piping in Ceilings. Pipe or tube is embedded in a metal lath and plaster ceiling. (19 mm) above the undersurface of the slab. Plaster is then applied to the lath to embed the tube. :o . V. (14 mm). 4.o:: o . (30 m) lengths. but higher temperatures drops can be used if applicable. If the slab is to be used without plaster finish." • ° : • ?".?• £ \ "•. Plaster is then applied to the metal lath. • -o'. 13 Metal Ceiling Panels Metallurgical^ BondedtoCopper Tubing ^METAL L A T H FINISHEO PLASTER C E I L I N G JOISTS Fig. the lath and heating coils are securely wired to the supporting members so that the lath is below. l £ SQUARE HEADER Fig. ' o ' : •[: o. Plastic. Otherwise. Coils are usually the sinuous type. (3 mm).6 mm) of cover below the tubes when they are installed below the lath. angle iron. Any supports used for positioning the heating coils should be nonabsorbent and inorganic. 2. the coil piping is installed as described for slabs resting on grade. The coils are constructed as sinuous-continuous pipe coils or arranged as header coils with the pipes spaced from 6 to 18 in. It is suggested that reinforcing steel. These include: (1) electric heating cables that may be embedded in concrete or plaster or laminated in drywall ceiling construction. Figure 18 shows the application of pipe coils in slabs resting on grade. the water supplied to the panels should not be Higher than 20 deg F (11°C) above the prevailing room temperature at that time and not in excess of 90 F (32°C). depending on the required output. ferrous and nonferrous pipe and tube are used in floor slabs that rest on grade. Insulation should be placed above the coils to reduce reverse loss. A warm-up and start-up period for concrete panels should be similar to that outlined for plaster panels. Water should be circulated at this temperature for about two days. Coils may be plastic. After paint and paper have been applied. A waterproofing layer is desirable to protect insulation and piping. cement or ceiling lath material. For plastered ceiling panels. with a minimum of 0. Figure 19 indicates one common type of construction. 19 Warm Air Floor Panel Construction ms-m •FOUNDATION Fig. In others.38 in. although some header or grid-type coils have been used in ceilings. (40 to 100 mm) of cover above the coils. ceiling panels are factory-assembled units furnished in standard lengths of about 75 to 1800 ft (25 to 550 m). Electrically Heated Ceilings Several different forms of electric resistance units are available for heating interior room surfaces. (9.5 to 9 in. an additional shorter warm-up period. construction codes may affect their position. Air-Heated Floors Several methods have been devised to warm interior room surfaces by circulating heated air through passages in the floor. 3. concrete block or similar materials should be used for support of coils. 17 Coils in Plaster Below Lath FOOTING WALL Fig. Although not as universally used as ceiling panels. pieces of pipe or stone or concrete mounds be used. No wood.8°Q per day to 140 F (60°C). water absorption. The coils are generally installed with 1.1-m) nonheating leads for connection at the thermostat or junction box. In some cases the heated air is recirculated in a closed system. brick.8. wall panels can be constructed by any of the methods outlined for ceilings. tube or channels built into the panel sections. 18 Coils in Floor Slab on Grade wood paneling and so forth.5 to 4 in. The outside diameter of the insulation covering is usually about 0. similar to first-time starting. Insulation is recommended to reduce the perimeter and reverse losses. the difference between heat supplied to the coil and net useful output to the heated room. is also recommended. a standard threecoat gypsum plastering specification14 is followed. or incorporation into.75 W per linear ft (9 W/m) and are supplied in capacities from 200 to 5000 W in roughly 200-W increments. (1 IS to 230 mm) on centers. During the air-drying and preliminary warm-up periods. (ISO to 450 mm) on centers. aging effects and chemical action with plaster. heat must not be applied to the panels for two weeks after all plastering work has been completed. Each cable unit is supplied with 7-ft (2. (2) prefabricated electric heating panels to be attached to room surfaces and (3) electrically heated fabrics or other materials for application to. there should be adequate ventilation to carry moisture from the panels. pipe or tube size and other factors. When the system is started for the first time. finished room surfaces. Standard cable assemblies are available for 120. The construction for piping embedded in floors depends on whether the floor is laid on grade or above grade. This insulation is normally a polyvinylchloride (PVC) covering which may have a nylon jacket. No paint or paper should be applied to the panels before these periods have been completed or while the panels are being operated. Compliance with applicable building codes is important. the heating cable may be sta- D-12 . all or a part of the air is passed through the room on its way back to the furnace to provide supplementary heating and ventilation. Where plastering is applied to pipe coils. 1. ferrous or nonferrous pipe or tube.208 and 240 V. Electric heating cables for embedded or laminated. Generally. Electric cables for panel heating have electrically insulated coverings resistant to medium temperature. Where the coils are embedded in structural load-supporting slabs above grade. Coils should be embedded completely and should not rest on an interface. the surface temperature of plaster panels should not exceed 120 F (49°C). These cable lengths cannot be altered in the field. with warm water piping. This can be accomplished by limiting the water temperature in the pipes or tubes in contact with the plaster to a maximum temperature of 140 F (60°C).11 Panel Heating and Cooling Systems LATH APPLIED' BELOW J O I S T S COILS W I R E T I E D TO L A T H OUTSIDE WALL ASPHALT IMPREGNATED INSULATION BOARD POURED SLAB CONCRETE EDGE CURBING PRECAST SLAB CONCRETE FLOOR SUPPORTS / AND DIRECTIONAL VANES -AIR PLENUM -CONCRETE -INSULATION 3 COAT FINISH HEATING COILS BELOW L A T H PLASTER -GRAVEL OR ROCK F I L L FINISHED PLASTER CEILING -EARTH Fig. The cable assemblies are normally rated at 2.12 in. To protect the plaster installation and to assure proper air drying. then increased at a rate of about 5 deg F (2. with coil pipes spaced from 4. ft2 (m2). The plaster is applied parallel to the heating cable. Heating cables or panels must be installed only in ceiling areas which are not covered by partitions. STAPLE SPACING 16 IN Fig. (64mm) clearance between adjacent cable runs must be left centered under each joist for nailing. and the two layers are held apart by the thickness of the heating cable. The spacing between adjacent runs of heating cable can be determined using Eq. With metal lath or other conducting surfaces. Where possible. It is essential that the space between the two layers of lath be completely filled with a noninsulating plaster or similar material.CHAPTER 8 8.4 MM) FROM TURN 3 IN. Cable runs that cross over the joist must be kept to a minimum. the system should not be energized and the range and rate of temperature change should be kept low by other heat sources or by ventilation until the plaster is thoroughly cured. it is permissible for a single run of isolated embedded cable to pass over a partition. 1984 Systems Handbook Electric heating cables are ordinarily installed with a 6-in. A 2-in. The purpose of this fill is to hold the cable firmly in place and to improve heat transfer between the cable and the finished ceiling. trim and ventilating or other openings in the ceiling. In drywall ceiling construction. An 8-in. (150-mm) nonhealing border around the periphery of the ceiling. For laminated drywall ceiling panels. While new plaster is drying. Failure to fill the space completely between the two layers of plasterboard may allow the cable to overheat in the resulting voids and may cause cable failure. C = length of cable. 20 Electric Heating Panel for Wet Plastered Celling D-13 EXTERIOR FINISH .:rJ IN.4 MM) INSULATION MINIMUM 1 . (76. A 2.75 W/ft (9 W/m). (200-mm) clearance must be provided between heating cables and the edges of the outlet or junction boxes used for surface-mounted lighting fixtures. (152 4 MMI CLEAR SPACE FROM WALL i FULL THICK INSULATION -• -irv STAPLE 6 IN. The National Electric Code requires that all general power and light wiring be run above the thermal insulation or at least 2 in. A„ = net panel heated area. the heating cable is always installed with the cable runs parallel to the joist. the LEAVE 8 IN. (152.2 MM> CLEARANCE BETWEEN BOX AND HEATING WIRE 6IN.(152.000 ohms measured to ground. a coat of plaster (brown or scratch coat) is applied to completely cover the metal lath or conducting surface before the cable is attached. Vermiculite or other insulating plaster causes cables to overheat and is contrary to code provisions. each cable is tested for continuity of circuit and for insulation resistance of at least 100. ft (m). rather than across the runs. (10): (10) s=UA„/C where s = cable spacing. cabinets or other obstructions.12 pled to gypsum board. The plaster fill should be applied according to manufacturer's specifications. (203.2 MM) FROM TURN AND ON RADIUS OF BEND (MAX. the heating cable is placed between two layers of gypsum board. The entire ceiling surface is finished with a covering of thermally noninsulating sand plaster about 0.S0 to 0. However. After fastening on the lath and applying the first plaster coat. The cable is stapled directly to the first (or upper) lath.7S in.5-in. (13 to 19 mm) thick or other approved noninsulating material applied according to manufacturer's specifications. Figure 20 shows details of ceiling cable installation practice for plastered construction. in. plasterboard or other thermally noninsulating fire-resistant ceiling lath. plaster lath or similar fire-resistant materials with rust-resistant staples. For cable having a watt density of 2. (SI mm) clearance must be provided from recessed lighting fixtures. (mm). these crossings should be in a straight line at one end of the room. (51 mm) above the heated ceiling surface. or that the wiring be derated. The predetermined cable spacing is maintained by daubs of cement. Some manufacturers recommend a minimum spacing of 2 in. (64-mm) clearance required under each joist for nailing in drywall applications occupies one-fourth of the ceiling area. 12 and 21. for drywall construction.5-in.18 to 9. twin. (10) is the net ceiling area available after deducting the area covered by the nonheating border. After the cable has been placed. The cable is installed on top of the first pour of concrete not closer than 2 in. 2. Preformed mats can be embedded in the concrete in a continuous pour. Therefore. Maximum operating temperatures vary from about 100 to about 300 F (38 to 149 °C).2 m) to 6 x 12 ft (1. (38 mm) and not more than 2-in. (51 mm) for drywall construction. Eq. Nonheating leads are connected and furnished as part of the panel. The mats are positioned in the area between expansion and/or construction joints and electrically connected to a junction box. (75-mm) thick and. This top layer must not be insulating concrete (see Fig. (10) contains a slight safety factor. 240 and 277 V service.75. (51 mm) from adjoining walls and partitions. 3. for simplicity. A cable assembly consists of the specified length of heating cable. in Eq.conductor and double cable. others must be installed as received. Some of the prefabricated panels described in the preceding section are also used for wall panel heating. The first pour should be at least 3-in. if the joists are 16 in. waterproof hot-cold junctions.1-m) cold sections. (2) lead and (3) tetrafiuoroethylene (Teflon). lighting fixtures.Panel Heating and Cooling Systems 8. Special anchor devices are available that are nailed to the first slab to hold the cable in position while the top layer is being poured. metallic or glass fiber mesh. are sometimes used for concrete floor heating systems. highly durable. These panels are available in sizes from 2 x 4 ft (0. At least 1 in. laminated conductive coatings or printed circuits. flexible heating cable composed of solid electric-resistance heating wire or wires surrounded by tightly compressed magnesium oxide electrical insulation and enclosed by a metal sheath. Net panel area. Electrically Heated Wall Panels Cable embedded in walls similar to ceiling construction is occasionally found in Europe. Remaining cable is then spread over the balance of the ceiling. steel and vinyl. such as those used for ceiling panels. 21 Electric Heating Cable in Concrete Slab variety of standard voltages.2 m) modular tee-bar ceilings. the cable can usually be stapled directly to the slab using hand-operated or powered stapling machines. (38 mm) for the first 2 ft (600 mm) from the cold wall. Because of possible damage from nails driven for hanging pictures or from building alteration. and a monolithic slab results. glass. They are constructed from a variety of materials such as gypsum board. Several standard Ml cable constructions are available. 7-ft (2. These mats usually consist of PVC-insulated heating cable woven in. Allow for circumvention of obstructions in the slab. Mineral-insulated (MI) heating cable is another effective method of slab heating. (400 mm) o. Consult the national and local codes for restrictions on the location of partitions.3 m 2 ) and with various watt densities. (Sl-mm) thick. with a bonding grout applied. must be multiplied by 0. UL-approved end fittings and connection leads. the net area.5 in. MI cable is a small-diameter. Many installations have a spaaing of 1. the finish slab should be poured within 24 hours of the first pour.5 in.5 in. A„. The cable is stapled to wood nailing strips fixed in the surface of the rough slab. Some panels can be cut to fit available space. such as single conductor. watt densities and lengths. Methods of fastening the cable to the concrete include: Electrically Heated Floors Electric heating cable assemblies. The slab is poured to within 1. The finish layer should be at least 1. depending on watt density. The 2. Custom-designed MI heating cable assemblies can be ordered for specific installations. Fig.c. should be insulating concrete to reduce downward heat loss. The surface is rough screeded. are finished as part of the ceiling. Rigid panels that are about 1-in. cabinets and other ceiling obstructions.6 x i . MI electric heating cable can be installed* in concrete slab using either one or two pours. where practical. Floor Heating Cable Installation When PVC-jacketed electric heating cable is used for floor heating. 208. (38 to 51 mm) of the finished level. For single-pour applications D-14 . the concrete slab is laid in two pourings. In lightweight or uncured concrete. in some cases.5 to 2 in.6 m). small lighting fixtures are usually ignored in determining net ceiling area. or attached to. and the mats placed in position. (25-mm) thick and weigh about 25 lb (11kg) each are available to fit standard 2 x 4 ft (0. there is no adhesion problem between the first and second pour. Panels may be either flush or surface mounted and. these assemblies must be carefully installed. For a proper bond between the layers. Since the first pour has not set. Prefabricated Electric Ceiling Panels A variety of prefabricated electric heating panels are available for either supplemental or full room heating. most codes in the United States prohibit such panels. Since. Since the possibility of cable damage during installation is greater for concrete floor slabs than for ceiling panels. Panel heating elements may be embedded conductors. Preformed mats are sometimes used for electric floor slab heating systems.6 x 1. all unnecessary traffic should be eliminated until the concrete covering has been poured and hardened. Always follow the installation instructions furnished by the manufacturer. (25 mm) of perimeter insulation should be installed as shown in Fig.8 x 3. MI cable is available in stock assemblies in a 1.13 minimum permissible spacing is 1. Different panels have rated inputs varying from 10 to 95 W/ft 2 (108 to 1023 W/m 2 ) for 120. Such mats are available as prefabricated assemblies in many sizes from 2 to 100 ft2 (6. 21). The final cap is applied immediately. Other outer-covering materials that are sometimes specified for electric floor heating cable include: (1) silicone rubber. lights and air grilles adjacent to or near electric panels. A„. A variety of contours can be developed by using heater wire attached to glass fiber mats. (38 mm) between adjacent runs. plaster of paris or tape. Thermal expansion of the ceiling panels must be considered. SHR = RSH . when a throttling valve modulates.4 to 2. this problem can be solved by using a combination series-parallel arrangement.6 m) wide at the outside wall can be designed for 235 F (113°Q surface temperature. Experience indicates that radiantly heated rooms are more comfortable under these conditions than when the thermostat is located on a back wall. Steel pipe should be the all-welded type. no threaded joints should be used for either pipe coils or mains. Piping should be designed to assure that water of the proper temperature and in sufficient quantity will be available to every grid or coil at all times.S fps (0. the water channel or lateral length should be greater than the header length.3MB to - ISO Fig. It is important to check with the latest issue of the National Electric Code and other applicable codes to obtain information on maximum panel watt density and other required criteria and parameters. a two-panel arrangement (Fig. design and installation of panel systems have certain requirements and techniques that should be recognized: . Reverse-return systems should be considered to minimize balancing problems.IF . Some types of plastic pipe also may be suitable where codes permit. and the radiant effect of the ceiling on the cover tends to alter the control point so that the thermostat controls 2 to 3 deg F (1 to 2°C) lower when the outdoor temperature is a minimum and the ceiling temperature is a maximum. 8. • Where possible. the dehumidification equipment is operating properly and building humidity is at design value. SUPPLY % i HC -*- i CO Vbvi R E T U II N - > * T i 0V2 T—i (el TWO-PIPE SYSTEM IM FOUR-PIPE SYSTEM Fig. because much of the panel heat transfer is lost to the return air system. For a given floor heating cable assembly. 7. Proper spacing between adjacent cable runs is maintained by using prepunched copper spacer strips nailed to the lower slab.6 or 0.32l'l 8 h X 0. 6. or for sinuous or serpentine flow. high velocity or high pressure drop devices or from pump and pipe vibrations must be avoided. 12. The panel area within 3 ft (1. the required cable spacing is determined from Eq. 23) can be used. Thus. Where coils are embedded in concrete or plaster. Design piping systems to accept thermal expansion adequately. 13. 2." b. Do not allow forces from piping expansion to be transmitted to ceiling panels. the design parameters are: a. there will be a rapid response. Noises from entrained air. Placing the thermostat on a side wall where it can see the outside wall and the warm ceiling should be considered. avoid automatic air venting devices over ceilings of occupied spaces.46 m/s) or higher] to prevent separated air from accumulating and causing air binding. Water velocities should be high enough [usually 1. The apparatus dew point of the cooling coils in the air distributing system should be designed for full capacity plus a 10 to 15% safety factor. 14. For two-layer applications the cable is laid on top of the bottom structural slab and embedded in the finish layer. or the last one or two rooms on the mains should have a bypass valve to maintain water flow in the main. the circulating water temperature should be maintained at room temperature until the air system is completely balanced.PC RTH where SHR = sensible heat ratio RSH = room sensible heat PC = panel cooling RTH = room total heat 3. uneconomical panel and a feeling of coolness at the outside wall. either the end of the main should have a fixed bypass. Individual ceiling panels can be connected for parallel flow using headers. If throttling valve control is used.6-m) border next to cold walls. If the laterals in a header grid are forced to run in a short direction.14 CHAPTER 8 the cable is fastened to the top of the reinforcing steel before the pour is started. c. As wilh any hydronic system.8. 11. F 750 I — i I 1! 1 300 Suggested Design Ceiling Surface Temperatures at Various Ceiling Heights panels into the plenum space above the ceiling are not recommended. but not less than ISO psig (1033 kPa). 10. The surface temperature of concrete or plaster panels is limited by construction. 5% antimony or capillary brazing alloys. With normal ceiling heights of 8 to 9 ft (2. both steel and copper pipe or tube are used widely in ceiling. close attention should be paid to the piping system design. (10). the panel surface temperature should be approximately as given in Fig. 1984 Systems Handbook i /\ / 40 1 | ! 1. 9. Changes in direction should be made by bending. Solder-joint fittings for copper tube should be used with a medium temperature solder of 95% tin. Higher watt densities [up to 25 W/ft 2 (269 W/m 2 )) are often specified for the 2-ft (0.7 m). Fittings and connections should be minimized. wall or floor panel construction. because most problems occur when the supply air is short on dehumidification capacity. Maintain adequate pressure in piping while pouring concrete. The normal thermostat cover reacts to the warm ceiling panel. 22 The application. Allow sufficient space above the ceiling for installation and connection of the piping that forms the radiant panel ceiling. Temperatures that are too low can result in an oversized. In circulating water systems. \y ' y / l IOO G E N E R A L DESIGN CONSIDERATIONS i 1 j 30 j 700 PANEL SURFACE TEMP. When the panel area for cooling is greater than the area required for heating. When the panel chilled water system is started. Excessively high temperatures over the occupied zone will cause ' the occupant to experience a "hot head effect. Copper tubing should be soft-drawn coils. 23 Split Panel Piping Arrangement for Two-Pipe and Four-Pipe Systems . In general. To avoid flow irregularities within a header-type grid. 22. d. 4.9 m) into the room. When selecting a ceiling panel surface temperature. The technique in item 7 above should be given priority. Ceiling system designs based on passing return air through the D-15 1 i i 20 1 CO NVERSIOR1 f ACTORS . If panels extend beyond 2 or 3 ft (0. 5.0 m) of the outside wall should have a heating capacity equal to or greater than 50% of the wall transmission load. panels less than 2 ft (0. Locate ceiling panels adjacent to the outside wall and as close as possible to the areas of maximum load. cable watt density and spacing should be such that floor panel watt density is not greater than 15 W/ft 2 (161 W/m 2 ). mean water temperature or watt density of an electric panel. All piping should be subjected to a hydrostatic test of at least three limes the working pressure. (2) size and location of the heating elements in the panel.Panel Heating and Cooling Systems 8. there is a considerable time lag between thermostat demand and heat delivery to the space. Frequently. 15. Indoor-outdoor thermostats are used to vary the floor temperature inversely with the outdoor temperature. Panel cooling systems require the following basic areas of temperature control: (1) exterior zones. Because the mean radiant temperature (MRT) within a panel heated space must increase as the heating load increases.8°C). To prevent condensation on the room side of cooling panels. In addition. In panel heating systems.5 deg F (0. as explained in Chapter 16. time clocks and current sensing devices are used on lighting feeders. Panels such as concrete slabs have large heat storage capacity and continue to emit heat long after the room thermostat has shut off the heating medium supply. D-16 . the panel water supply temperature should be maintained at least 1 deg F (0. since a large part of the heat must first be stored in the thermally heavy radiant surface. Therefore. particularly in applications where there is a high lighting load. the air temperature during this increase should be lowered I or 2 deg F (0. the ratio of outdoor temperature to slab temperature is 70:15 (39:8). resulting in a very slow reduction of the space temperature at night and a correspondingly slow pickup in the morning. Window pane thermocouples have been used to schedule water temperatures in panels under a window sill. Determine the available area for panels in each room. 17. (3) insulation on the reverse side and edge of the panel. to prevent condensation on window surfaces. very little fuel savings can be expected even with light panels unless the lowered temperature is maintained for long periods. In one application. Selection of summer design room dewpoint below 50 F (10°Q generally is not economical. The procedure is summarized as follows: 1. This minimum difference is recommended to allow for the normal drift of temperature controls for the water and air systems. However. as in kindergarten floors. The ambient sensing thermostat controls the comfort level. enabling a conventional room thermostat to be used. The temperature control of the interior air and panel water supply should not be functions of the outdoor weather. As with all systems. each exterior corner zone and similarly-loaded face zone should be treated as a separate subzone. Photoelectric cells can be used to divert cold water into a peripheral ceiling panel. hygroscopic chemical-type dewpoint controllers are required at the central apparatus and at various zones to monitor dehumidification. or approximately 5:1. If the main cooling coil is six rows or more. and also to provide a factor of safety for temporary increase in space humidity. Determine the required panel surface temperature. 22. When chilled water is used. and additional sensible cooling is necessary to cool the dehumidified air to the required system supply air temperature. Tests on a metal ceiling panel demonstrated the speed of response to be comparable to that of conventional environmental systems. or a long warm-up period should be provided. may respond to changes in demand quickly enough for moderately satisfactory results from lowered night temperatures. The remote sensing thermostat in the slab acts as a limit switch to control maximum surface temperatures allowed on the slab. 19. heat exchange or using the water leaving the dehumidifier is the major consideration in preventing condensation. This inertia will cause uncomfortable variations in space conditions unless controls are provided to detect load changes early. 21. For eeneral information on automatic controls.15 Panel HC (heating and cooling) is supplied with hot or chilled water year-round.6°C) above the room design dewpoint temperature.) is held to a maximum of 80 to 85 F (26 to 29°Q. the temperature of the heating medium supplied to the panel surface should be varied in accordance with outdoor temperature. PANEL HEATING SYSTEM DESIGN Design Steps Panel design requires specifying the following: (1) panel area. the supply air dew point should be reduced during extremely cold weather according to the type of glazing installed. Automatic controls for panel heating differ from those for convective heating because of the thermal inertia characteristics of the panel heating surface and the increase in the mean radiant temperature within the space under increasing loads for panel heating. when used as a primary heating system. However. Panel cooling systems have also been zoned to provide individual temperature control in exterior offices. the dewpoint of the air leaving will approach the temperature of the water leaving. (2) areas under exposed roofs to compensate for transmission and solar loads and (3) control of each typical interior zone to compensate for internal loads. lowered night temperatures will produce unsatisfactory results with heavy panels such as concrete floors. or for corner rooms with large glass areas on both walls. Calculate the hourly rate of heat loss for each room. These panels cannot respond to a quick increase or decrease in heating demand within the relatively short time required. (4) required input to panel and (5) temperature of the heating elements. If the heat loss of the building is calculated for 70 to 0 F (21 to -18°C). The most frequently applied method of dehumidification utilizes cooling coils. The normal thermostat drift is usually adequate compensation for the slightly lower temperatures desirable during winter weather. Calculate the required unit panel output. If reduced nonoccupancy temperatures are employed. See Chapter 31 for further information on automatic controls. 4. 20. such as plaster or metal ceilings and walls. Because air quantities are generally small.. and both panels are used for cooling. refer to Chapter 31. it is possible to avoid compromising indoor conditions with a panel cooling system throughout the year. A manual boiler bypass or other means of reducing the water temperature may be necessary to prevent new panels from drying out too rapidly (see Embedded Piping for Ceiling Panels).56°C) increase in the slab temperature. 16. Electric Heating Slab Controls. A time clock is used to control each heating zone if off-peak slab heating is desirable. In ordinary structures with normal infiltration • loads. When chemical dehumidification is used. Lightweight panels. For supplementary slab heating. Control of the interior zones is best accomplished by devices that reflect the actual presence of the internal load elements. When cooled ceiling panels are used with a variable air volume (VAV) system. 2. the surface of a floor slab {!. the controls function to activate panel CO (cooling only).15 However. but should be wired in series with a slabsensing thermostat. thermostatic control devices sensing air temperature should not be used to control the slab temperature. many of the control principles for hot water heating systems described in Chapters IS and 16 also apply to panel heating. cooling tower water is used to remove heat from the chemical drying process. particularly after extended down periods. It is imperative to dry out the building space before starting the panel water system. 18. Such delayed starting action can be controlled manually or by device. Panels are suitable for control systems which are scheduled by elements that sense solar and weather changes before these changes affect the space temperature. which might damage the panels if controls failed.5 to l°C)to maintain comfort. a remote sensing thermostat in the slab is commonly used to tune in the desired comfort level. the air supply rate should be near maximum volume to assure adequate dehumidification before the cooling ceiling panels are activated. For optimum results. This drift should be limited to result in a room temperature change of not more than 1. such as weekends. This means that a 5 deg F (2. The cooling water leaving the dehumidifier can then be used for the panel water circuit without danger of condensation during normal operation. Controls (Heating). Several chemical dehumidification methods are available to control latent and sensible loads separately. it is not advisable to use volume control in any part of the system. Controlling the panel water circuit temperature by mixing. the required reduction in air temperature is small. some means of providing a higher-than-normal rate of heat input for rapid warm-up is necessary. For comfort heating applications. In general. An ambient sensing thermostat is used to vary the ratio between outdoor and slab temperatures. and the floor temperature range is held from 70 to 85 F (21 to 29°C) with a remote sensing thermostat.g. for embedded pipe panels precautions must be taken to prevent the introduction of excessively hot water. With the apparatus arranged to supply air of appropriate apparatus dewpoint at all times. 3. e. Controls (Cooling).8°C) drop in outdoor temperature requires a 1 deg F (0. picking up the winter solar load on a south zone. Other considerations are listed in items 12 and 14. 6x2. Select the means of heating the panel and the size and location of the healing elements. find the panel surface temperature needed to yield the required heat output to each room. design room air temperature. Btu/h-ft2 (W/m 2 ). F (°C). that is. overall coefficient of heat transfer for the given construction between room air and the point tb. 8.8 = 62) 8000/306 = 26.4 = 82) Step 3. ft (m) represents linear ft (m) of exposed slab perimeter and deg F (°Q represents temperature difference between concrete surface and outdoor air. The result is the minimum heat output per square foot (square metre) of panel that will satisfy the requirements of the room. upward heat flow from panel. which simulate various conditions of construction and outdoor temperature. 8 for Room A is illustrated in Fig.3 = 150) 2500/127 = 19. Keep floor temperatures at or below 85 F(29°C).5x6. total thermal resistance of panel to upward heat flow. Letter Symbols for Examples of Design Methods The following design examples use the letter symbols shown below. outdoor design air temperature.42 W/m2 • °C). F (°Q. (Both. Procedure for Metal Ceiling Panels Design procedures for metal ceiling panels in a heating application are included in the example given in the section Design Procedure for Panel Cooling Systems. 7. F(°C). Panel Surface Temperature From Fig.28) from room C to outdoor air. F (°Q. Step I.4°C).3) RoomB 127(11. Room dimensions and calculated heat losses are as follows: Room Room A RoomB RoomC Dimensions ft(m) 11x12x8(3. Btu/h«ft2-F (W/m2«°C). '* = air temperature above or below panel at point to which U. Band C are as follows: D-17 . the effect of each assumption or choice on comfort should be considered carefully. The procedure is applicable within the following range: Outdoor design conditions: Temperatures as low as -30 F (-34. qd Btu/h-ft2(W«m2) 6300/132 = 47.7 ( 730/11.8°C). Select insulation for the reverse side and edge of panel. c. 8.7 m)." The procedure for designing a plaster ceiling panel is illustrated by Example 1. ft2»F»h/Btu (m2*°C/W). ft (m). Plaster Ceiling Panels. Btu/h* ft (W/m ). (Btu/h«ffF (W/m«°C). Room-scale tests. A panel designed by this procedure will maintain the desired room air temperature for the selected outdoor conditions. design mean water temperature (selected for each zone). mean water temperature. /> = length of exposed edge of slab. deg F (°C) represents temperature difference between panel surface and air.16 CHAPTER 8 5. or C 2 is taken.1 and 3. 2 2 Qd~ downward heat flow from panel.) Conventional interior finishes and furnishings. F (°Q.4"C). Determine panel heat loss and required input to the panel. Room dimensions: Rooms having normal proportions. inlet water temperature. Air changes: No more than two air changes per hour.8. Room air temperature is the selected criterion of comfort. Determine the other temperatures that are required or developed.4) Heat Loss Btu/h(W) 6300(1850) 2500(730) 8000(2300) Step 2. ceiling height between 7 and 12 ft (2.' . panel area.6x2. The panel that requires the highest output per square foot (square metre) will generally control the design.°C/W). '/> = panel surface temperature (exposed surface). = coefficient of downward and edgewise heat loss of exposed slab.7(1850/12. Room construction: Any type of wall construction and any amount of glass area. 2. F (°Q. Heat Loss Calculate the heat loss of each room by methods outlined in Chapter 25 of the 1981 FUNDAMENTALS VOLUME.25 Btu/h • ft2 • F (1.to-air U value of 0. they represent a single zone. 9. thermal resistance of bare concrete panel to downward heat flow. C/W).0 C/W). The following general rules should be followed: 1. but do not include any heat loss through the area covered by the panel. In the design steps. 6. Available Panel Area ft2(m2) Room A 132(12. A. ft2»F«h/Btu (m 2 '°C/W).05 (0. because the temperature of the fluid in the system must be high enough to produce the required output from that panel. thermal resistance of panel to downward heat flow. Design the system for heating the panels according to conventional practice.1 (2300/28. 8A. Place panels near the cold areas where the heat losses occur. outlet water temperature.8) RoomC 306(28. F (°Q. 'dc = thermal resistance of material between the underside of the concrete slab and the ceiling surface below.4) 11x12x8(3. Room air temperature: 70 to 76 F (21. C|. Required Panel Output Divide the heat loss of each room by the maximum ceiling area in the room that can be used as a heating panel. have shown that this near-equality of the two temperatures normally prevails. Btu/h«ft2«F (W/m2«°C). coefficient of heat transfer from the upper surface of the concrete slab which forms the ceiling panel to air above panel at point tbi Btu/h-ft 2 -F (W/m 2 -°C).3x3. using the output determined in Step 2 and the design room air temperature. Btu/h«ft2 (W/m 2 ). have an effect on comfort. apportioned downward and edgewise heat flow from Ide = panel. thermal resistance of floor covering. The ceilings of rooms A and B have floors above them with the space heated to 72 F (22.3x2.3x3.= coefficient of heat transfer from lower surface of concrete slab to air below the panel at point tb.B and C. The ceiling of room C has insulation in the joist spaces and an uninsulated attic space with a combined (/value of 0. ft 2 «F-h/Btu(m 2 .4) Required Panel Output. ft2 (m 2 ). water-heated panels. however. F ("Qsurface temperature of top of concrete slab. The procedures are based primarily on data obtained at the former ASHRAE Research Laboratory. Do not use high temperature ceiling panels in very low ceilings. The values of tp determined for Rooms A.2°Q air temperature when the outdoor air temperature is 0 F (-17. Example 1: Three rooms. have a common water supply temperature. deg F (°Q represents temperature difference between panel surface and air. This use of Fig. ft 2 «F«h/Btu(m 2 . c. 1984 Systems Handbook Warm Water Panels—Embedded Pipe This section presents a simplified procedure for the thermal design of embedded pipe. maximum water temperature permissible for a given construction. '•</ = total ft 2 -F'h/Btu(m 2 .4) 15x21x8(4. F(°C). They are maintained at 72 F (22. F 'dmw (°Q.1 to 24. r. ft2«F-h/Btu (m2«°C/W).2°Q and an air. glass and floor or ceiling does not differ greatly from room air temperature. thermal resistance of bare concrete slab to upward heat flow. The design procedure is restricted to situations in which the area-weighted average temperature of unheated surfaces of walls. The required panel output is given in the following table. 3. Design Mean Water Temperature Select a single design mean water temperature (jtjmw) f° r each group of rooms comprising a zone.-T •*^r-zZ. 8B.6°Q and Room C fmw = 119 F (48. In Example I.l lOO '20 140 HMP MINUS AIR 7IMP ABOVt TANT I.35(0. If fm„ + ((/.„ = 15 deg F (8. 8. Provide supplementary heating.3°C) and tmax = 140 F (60°C).5/47. This can be accomplished by either or both: 1.17 I CONVERSION FACTORS: • C . qu ±U(tp-1/. If tmw + l(f(.I F . 8 I . tm„ + 0. 24. 24A.3°Q.Panel Heating and Cooling Systems 8. 8 .5 F (43. and panel resistance (rd). 8 is used to find qd for Room B in this step. B and C as determined from Fig. Step 7. find the panel output (qd) for design mean water temperature Crfm„).„.8 W/m : -Btu/h-ft. this mean water temperature (tmw) is an acceptable design mean water temperature ltdmw). If the required mean water temperature is still too high.t0) = 138. choosing the highest mean water temperature (rm„) of the group subject to: a.I°Q Step 4. room air temperature (/„).0(18. 6 8 ^ (U factor) ct i i Fig.90(0. o CONVERSION FACTORS: C . (mm) 4. 131 F (55°C) can be used as the design mean water temperature tjmK.. Reducing the tube spacing. 8 is illustrated in Fig.). 24 are those for the entire structure of which the panel is a part and are taken from room air (temperature. 2. 24 is illustrated for Room A in Fig. 24 1. This use of Fig. 8C Determination of Design Panel Output for Room B from Fig. ROOM A B C SPACING in.3°C) RoomC t„ = 97F(36.ftJ• F • h/Btu X 0. Design Panel Output From Fig.5 2§S. f DEC Fig. go back to Step 5 and select a panel construction having a lower panel resistance (r d ).The panel output determined in Step 2 for Room A is therefore the design panel output Step 6. VP Ub.r„)/2] is greater than !„. 8B Determination of Mean Water Temperature for Room A from Fig.9/62) = 0.5 F (59.„. -C.5 (/.16) 5/26..8 t W/m"> Btu/h It-' X3.42) 0. The use of Fig.F • h/Btu X 0. assume t. find the resistance of each panel (rd) using heat flow ratios calculated from the flow rates determined in Steps 2 and 4. and the room air temperature (/„).deg F/1.Btu/h • ft-" X 3./ 0 )/2] is equal to or less than /„. 8 are as follows: Room A tmw = 131 F (55°C). 1 S W/nv C » Btu/h ft-" • F X 5 . Also in Table 3. SUWAt.1 (15.1°C) and is less than /„„„. 2.15 m'-'C/W-. . . ta) to air above the panel (temperature. For that reason..f t ! .1S tn!-'C/W.176 Ir value) 1 t 1- Fig. Then since tmvl = 131 F (55°C).5(114) 9 (229) 9 (229) HEAT FLOW RATIO RESISTANCE. The upward heat flow must be determined to obtain the downward panel resistance and to select the proper size of boiler. interpolating as required. Step 8.4/150) = 0. 8. find the mean water temperature (tmw) from Fig.5(39.Proceed to Step 8. |F-321/1. 8 CANU.8 W / m ! • Btu/h • I t ' X 3 .25(1.8) Step S. 24A Determination of Upward Heat Flow for Room A from Fig.. 24 Upward Heat Flow from Plaster Ceiling Panel Room A /. Room B tmw = 110. 8C shows how Fig. Reduce the heat loss of the room.m 2 ) 12. Thus._ 32.o i CONVERSION FACTORS: "C=|F-32)/1. The values of qu for Rooms A. b.9) 5.6°C) RoomB t„ = 92F(33.176 (rvalue) 7072> 4' Fig.X3.15 1 Fig.321/1. Upward Heat Flow Determine the upward heai flow from each panel using Fig.30 0. The mean water temperature (tmw) found for Room A in Step 6 was used as the design mean water temperature Udmw). / ft ).19 0. Values of tmw for Rooms A.85 (0. Decreasing the upward heat flow (qa) by increasing the insulation above the panel. = 114F(45.15) I'll. B and C as determined* from Fig. apply either or both of the following: D-18 __477_ H l=Oj/ "*Z? ^^"t>4 CONVERSION FACTORS: .7 (18.C .4) 6. . Mean Water Temperature For the required panel output (qd) found in Step 2. Note that the U values shown on Fig.8 WAtr .7 (39. choosing the closer spacings when higher outputs are required. the panel resistance (rd) found in Step 5.8/82) = 0.25(1.0(15. 24 are as follows: Room U A B C 0...26 0.42) 0. U values may be obtained from Chapter 23 in the 1981 FUNDAMENTALS VOLUME. rd ft2«F«h/Btu(m*«<'C/W) 12. Downward Panel Resistance Assume tentative pipe or tube spacings for the panel construction to be used from those listed in Table 3.06) 6/19.05 (0.Fig.28) tp-'b degFfQ 42(23) 20(11) 97 (54) Btu/h'ft 2 (W. 8A Determination of Panel Surface Temperature for Room A from Fig. 7(150. Step 6. Design Panel Output From Eq. If the design panel output is different from the panel output (qd) used in Steps 3 and 4. and the air temperature of the space above. These steps are identical to the corresponding steps for plaster ceiling panels. Mean Water Temperature and Upward Heat Flow a.5f(29..5f(17. Design panel outputs for Rooms B and C can be found from Fig. the heat flow upward (qu) should be redetermined. Upward Heat Flow Estimate a. The panel that requires the highest output per ft2(mJ) will generally control the design. use: <7«=C. Heat Loss Calculate the heat loss of each room. Mean Water Temperature For the required panel output found in Step 2. using Table 2 to find both resistances (r„. the rooms should be subdivided into areas having somewhat similar heat requirements. 9. Step 3.3) 47. Steps 4 and 5. 9.7(150) 32.Qu + Qd) A B C 12.10. Follow the procedure for Plaster Panels. using the panel surface temperature (tp) found in Step 3.3) 77 ( 7. 8 using the design mean water temperature (tdmw).2 and 3. and Fig. Panel Surface Temperature Assume a trial panel surface temperature and determine the resulting heat output from Fig.1°C) between the water inlet and outlet of the panel (see applicable chapters on hot water systems).9) 5. Step 3. 9. I f the upper surface of the slab is not exposed. Step S. (See the section on Concrete Floor Panels— Intermediate Slab.5 (39. These steps are identical to the corresponding steps: for plaster ceiling panels. Step 6. II and 12.3) (unchanged from Step 2) 77 ( 7. 9. separate the interior areas requiring little or no heat input from the exterior areas directly influenced by outdoor weather conditions.).6 to 11. i. Step 8. Concrete Floor Panels (Slab-On-Grade) Step I. The result is the minimum heat output per sq ft (square meter) of panel that will satisfy the room requirements. ft2 (m*) 132 (12.8. Step 4. Design Panel Area Divide the room heat loss found in Step 1 by the design panel output found in Step 8. (t„-tb) (11) Step 5. Assume successive trial panel surface temperatures until finding the temperature at which the combined heat transfer from the panel equals the output determined in Step 2. but a design can be approximated using the equations of heat transfer from walls together with the thermal resistance properties of plaster ceilings from Table 3.) The procedure for hot water plaster ceiling panels cannot be wholly applied to concrete ceiling panels because some of the simplifying assumptions regarding the upward heat flow from plaster panels are not valid for concrete panels.0(104) Step 9.5(102) 33.e. A design graph has not been prepared for wall panels.5 (102. Step II. b. 2 and 4. qj Btu/h»ft2 (W»m2) 47. Concrete Ceiling Panels. The net Btu (kj) rating of the boiler should equal or exceed the total output of all panels plus any other loads on the boiler.4) 33. Floor panel surface temperatures exceeding about 85 F (29°Q are D-19 . Treat each area as a separate room for design purposes. but do not include any heat loss through the area covered by the panel.5) Btu/h(W) 7946(2328) 3234 ( 948) 9317(2730) t Redetermined. Panel Surface Temperature From Rg.2) 242(22. Steps 9. The procedure for plaster ceiling panels is used as a guide. the panel resistance found in Step 5. Design Mean Water Temperature Follow the procedure for plaster ceiling panels. Both types of panels have heat outputs in two directions in amounts determined by the thermal resistance and the temperature difference in each direction. Add to the thermal resistance of the slab to upward heat flow (r u ). Concrete ceiling panels are distinguished from concrete floor panels in intermediate floors by the position of the tubes in the concrete slabs (see Tables 1 and 2). If very large rooms are involved. The mean water temperature (tmw) found and the air temperature of the space above the panel are the other two factors to be used. Required Panel Output Divide the heal loss of each room by the maximum floor area in the room that can be used as a heating panel. Total Panel Output Add the heal flow upward (?„) to the design panel output (qd) and multiply by the design panel area to obtain the total panel output. A.11 and 12.0(104.) of any material between the upper surface of the slab and the space above to obtain the resistance (r„) to be used in Fig. find the panel surface temperature needed to yield the required heat output to each room by using the output determined in Step 2 and the design room air temperature.5) Step 10. (12). Design Mean Water Temperature Select the highest mean water temperature (imw) as the design mean water temperature (/. Room Qu Qd Ap Total Panel Output Ap (. b.2) 242(22. The effect of these outputs on space heating requirements and the occupants' comfort should always be considered. Steps I and 2./««. find the panel output for design mean water temperature tdmw by successive trials. Fluid Circuit Design the fluid circuit (panel piping and mains) for a temperature drop of 10 to 20 deg F (5. calculate the required mean water temperature as follows: tmw='p+rdlQd) (12) Step 7. Step 6. and rd). If the upper surface of the slab is exposed to form a floor. Plaster Wall Panels.9. These steps are identical to the corresponding steps for plaster ceiling panels. Room A B C Design Panel Area. Upward and Downward Panel Resistance 1984 Systems Handbook Follow the procedure for Plaster Panels. Find the heat flow upward from the panel (qu) from Fig. 2 and 4 as explained in the accompanying section of the text. Room A B C Design Panel Output. Step 12. Step 7. because the temperature of the water in the system must be high enough to produce the required output from that panel. Step 7.3) 32.10. Steps 8.18 CHAPTER 8 for this room. and the room air temperature. the resistance to heat flow (/•„. The plaster ceiling panel procedure must be modified as follows: Steps 1. These steps are identical to the steps for plaster ceiling panels. Step 2. Boiler Size Size the boiler according to the method in Chapter 24 of the 1983 EQUIPMENT VOLUME. These steps are identical to the corresponding steps for plaster ceiling panels.0) 132(12.4) 9. And the heat flow upward from Fig. If the underside of the concrete slab is not exposed. Step 9.) and outlet (t0) of the panel (see Chapters 15 and 16).05 Btu/h«ft 2 «F (0. 25 or Eq.. These steps are identical to the corresponding steps for slab-ongrade construction.2 and 3. c. Design Panel Output From Fig. Use the heat loss per ft2 (m2) of panel area from Step 2 and the upward thermal resistance from Step 3. Determine the downward and edgewise heat loss coefficient C3 from Fig. find the heat output downward (qd) from Fig. Upward and Downward Panel Resistance Follow the procedure for slab-on-grade construction. find the design panel surface temperature. Step 4.5 w7m2-°C). Estimate the heat flow downward as follows: 1. Ru. tofindthe total resistance to downward heat flow (rd) to be used in Fig. masonry slab construction: Step 1. Also from Fig. b. U = 0. the apportioned downward and edgewise heat flow (q^ should be redetermined.413 Btu/h per watt [watt output/watt input (W 0 /Wj)] and the nearest standard cable assembly is selected.5 F (46. find the thermal resistance (rm) of the slab of each panel. using heat flow ratios (qu/Qde) calculated from the flow rates determined in Steps 2 and 4. 9.4°C) (see Step 3) in a room. Safety factors are usually omitted in the design of electric heating systems. Step 6. Step 5. 8 to find the downward heat flow (qd). and the room air temperature (ta).4 Btu/h«ft2«F (2.19 not recommended since they wilt probably cause foot discomfort. Step 3. Electric Ceiling Panels on Masonry Slab (Intermediate Floors) The following steps give the design procedure for ceiling heating panels that consist of electric heating cable embedded in plaster applied to the lower surface of intermediate floor. These steps are identical to the corresponding steps. If the design panel output is appreciably different from the panel output (qu) used in Steps 3 and 4. Calculate the heat loss of each room. the design heat loss is most commonly calculated in the normal manner. Do not include the surface film resistance. Steps 7. Add to the slab resistance to downward heat flow (r^) the resistance to heat flow (r^) of any material between the underside of the slab and the ceiling surface. Design Mean Water Temperature Select a single design mean water temperature (tdmw) for each group of rooms comprising a zone. Steps 1.) by adding to the panel surface temperature the temperature difference caused by the thermal resistance of the floor covering. interpolating as required.Panel Heating and Cooling Systems 8. Use either Fig.97 Btu/h-ft 2 -F(5. 2. In the example. Determine the temperature of the surface of the concrete slab (/. 9.) of the group. Insulation with a conductance of 0. Step 6. 8. 25 is used: D-20 .1°Q between the water inlet (/. The net Btu (kJ) rating of the boiler should equal or exceed the total output of all panels plus any other loads on the boiler.6 m) below the slab results in a slab downward and edgewise heat loss coefficient of 0. If it exceeds 85 F (29. StepS. and panel resistance (r u ). To the slab resistance (r^). The design heat loss is then converted to watts in the ratio of 3. Step 4. Step 2. Boiler Size Select the boiler size according to the method in Chapter 24 of the 1983 EQUIPMENT VOLUME. Insulated Electric Ceiling Panels Electrically heated ceiling panels can be designed using the equations and curves previously presented in this chapter. Total Panel Output Follow the procedure for slab-on-grade construction. Step 11. Determine the combined thermal resistance. However. Design Panel Area Divide the room heat loss found in Step I by the design panel output found in Step 8. repeat Steps 5 and 6 using the calculated value. 9. 116. Step 4. (14) Step 5. Mean Water Temperature For the required panel output (qu) found in Step 2. If the underside of the concrete slab is exposed to form a ceiling. Follow the procedure for slab-on-grade construction. Cable length is ordinarily published as part of the manufacturers* rating data. These steps are identical to the corresponding steps for slab-ongrade construction. If Fig. If the heat flow downward (qd) differs appreciably from the estimate made in Step 4. Apportion the downward and edgewise heat loss uniformly across the panel as follows: Ode=PxCi{ts -toaVA. Fluid Circuit Design the fluid circuit (panel piping and mains) for a temperature drop of 10 to 20 deg F (5. Step 4) to obtain the panel resistance to upward heat flow (/•„). add the resistance of the floor covering (r^. use the equation: Qd = C2 Us ~ lb) C5) Step 5. Concrete Floor Panels (Intermediate Slab). Step 10. Mean Water Temperature and Downward Heat Flow a. Step 7.. Step 12.9°Q is used.28 W/m 2 '°C). Use Fig. Step 10. Total Panel Output Add the apportioned downward and edgewise heat flow (<7rfc) to the design panel output (qu) and multiply by the design panel area to obtain the total panel output. the panel resistance (/•„) found in Step 5. room air temperature (r a ). Downward and Edgewise Heat Flow a. 11 for the insulation to be used. The water temperature (tmw) found above and the air temperature (/6) of the space below the ceiling are the other two factors to be used. Down ward Heat Flow Estimate a.6 to 11. part a. the reverse heat loss is considered negligible. Step 4. Therefore. Determine the required watts input per ft2 (m2) of panel area. b. (16) and (17) as follows: a. Cable spacing is calculated from Eq. Step 6. If the required heat output cannot be obtained from an 85 F (29°Q floor panel. using Table 1 tofindboth resistances (rm and r^). substituting heat flow downward (qd) for the apportioned downward and edgewise heat flow {q^). if any: 's = 'p + (lu * rM) (13) The value of rue for various floor coverings is given in Table 5. find the panel output (qu) for design mean water temperature (tdmw). The procedure for designing a hot water concrete floor panel of intermediate slab type is given in the following steps. for slab-ongrade construction. Steps 11 and 12. Step 10. choosing the highest mean water temperature (/„. Follow procedure for slab-on-grade construction. heat losses should be reduced or supplementary heating should be provided.8 and 9. choosing closer spacings and larger pipe or tube when higher heat outputs are required. but do not include any heat loss through the area covered by the panel. of the slab material or materials above the heating cable. Upward Panel Resistance Assume a tentative pipe or tube size and a spacing for each panel. using the design heat transfer coefficients as given in the 1981 FUNDAMENTALS VOLUME. using the slab surface temperature (ts) found in Step 3 as the panel surface temperature (tp) and the air temperature of the space below.„. Divide the heat loss of each room by the net ceiling area that can be used as heating panel. From Table 1. for ceiling panels which have the thermal insulation recommended for electric heating.3 W/m 2 -°Q extending 2 ft (0. find the mean water temperature Umw) from Fig. go back to Step 5 and choose a wider spacing or smaller pipe or tube for that room. (10). b. 8. CHAPTER 8 8. Step 7. The heat output to the space above the slab equals the difference between the heat energy input and the ceiling panel output. Select from manufacturers' ratings the standard cable assembly that most closely matches the required total watts. The procedure can be used for small deviations from these conditions. (15) and (16) were derived using Eq. and the remainder is lost through the slab for a distance of approximately 3 ft (0. A slab of infinite length and width. Btu/h-ft 2 (W/m 2 ). W. 25 Ceiling Panel Output for Intermediate Floor.176 Ir value) IS 20 HEAT INPUT. Step 8.. (3)..3211. (Wa/Wi).854 Ru .h - F h/BuiX 0. also ceiling panel output. Steady-state heat flow. (16) and (17) are used: Determine the value of m from Eq. = qc/m m = 1.9°C) both above and below the slab. Electric heating cable embedded in a material having a thermal resistance below the cable of 0.06). With m and the room heat loss per ft2 (m2) of panel area from Step 2. Ambient temperature of 75 F (23. it is important to thermally isolate the heated slab from the adjacent floor.8 nr B ( I X 0. Watt input can be determined from Eq. * 75 F ambient and 75 AUST above and below slab.0929 : m: ty w .0. (16). drop to the horizontal scale and find the necessary Wj/ft2 (Wj/m 2 ). * . •75 F ambient and 75 AUST above and below slab. For other conditions. ft2-F«h/Btu (m2-°CAV). (17) can only be used for values of Ru between 0. (16) and (17). b. 26 20 HEAT INPUT. This procedure is based on the following conditions: 1. Blu/h per W. 26. floor coverings or other means.9m) from the exposed edge. This edge loss is proportional to: (1) the difference between indoor and outdoor temperatures.02). A large amount of the heat is lost through the exposed edge of a floor.0 (1.. WATTS PER SO FT. Upper Surface Temperature of Intermediate Floor. Multiply the required watts input per ft2 (m2) by the net panel area available to obtain the total watts required for the room. 85 /.4°C). WATTS PER SQ F*.0 (0. 5.013 Ru3 (16) (17) [0. Fig. 3.20 1984 Systems Handbook 1 / 1 / ' 80 f/ 5* / 4 1 "J ezr i / / / j / . / . A portion of this loss is directly through the concrete to the air.09) S * „ £ 6. m = ceiling panel output factor. exposed edge and earth.' H^KUti-O where D-21 (18) .9°C) both above and below the slab. Masonry Slab Structure with Embedded Cable Ceiling Panel* Ru = the combined or equivalent thermal resistance of the slab material above the heating cable. in Btu/h (W) per ft (m) of exposed edge. 2. qc => design heat loss per ft2 (m2) of panel area./ / ' CO 4 f / i CON\/ERSION FACTORS: C» f .175 Rul + 0. Determine the proper spacing between adjacent cable runs using Eq. Figures 25 and 26 and Eq. Masonry Slab Construction* Enter the chart on the vertical scale at the required ceiling panel heating output (from Step 2).06)] where W. 4. Step 6. (17). c.9-m) border along the exposed edge. This resistance does not include the surface film resistance. (7) and (8). AUST of 75 F (23. the basic panel heat output equations should be used. Supplement the panel heat output with another form of electric heat. If Eq. If the resulting floor temperature exceeds 85 F (29.5 (0. Note: Eq.W. calculate watts input from Eq. = watts input per ft2 (m2) of panel area. Ru does not include the surface film resistance. Step 5. This may be done by using insulating concrete. one of the following steps should be taken: a. Reduce the room heat loss by increasing the building insulation. (10). At the intersection of the required heat output and the curve of thermal resistance for the particular construction. (2) the length of the floor perimeter adjacent to the exposed edge and (3) a heat loss factor (K).. 5 10 IS Fig. Increase the insulating value of the slab above the insulation.5 and 6.344 + 0. deg F (°C) difference between indoor and outdoor temperatures.09 to 1.12 (0. Electric Floor Slab Heating In electric floor slab heating. b. based on the amount of heat loss through the floor area included within a 3-ft (0. Determine the approximate floor surface temperature from Fig. 500(1026) 4400(1319) 14. asphalt or tile).t.5 W/m 2 ) or (conservatively) 1 W/ft 2 (11 W/m 2 ). The remainder of the heat generated in the inner floor will be stored and. B and C. 6. the three rooms. 13. 3.5 W/ft 2 (5. F 0EG. hd = downward heat loss. under certain conditions of temperature difference. 27 Metal Ceiling Panel Design Graph Showing Panel Surface Temperature and Mean Water Temperature Difference vs. ft (m). Warm Air Panels The first three steps in the design of warm air panels are the same as those outlined for hot water panels and the same per- ^ -yd? X s* ' • & - > * &^-— jfc> CONVERSION FACTORS: -C-degF/1.7x2.ta) (20) where h0 = heat output.9 m) of the edge and (2) the heat loss through the remaining (inner) floor area. Establish minimum supply air quantity. = heal input. The heat transmission coefficient U has been determined17 to be 0. K = h. 2. 1 W/ft 2 (11 W/m 2 ). Fig. W. -h„ = U(ts . 12.4) 15x21x8(4. 6 bate) 25 S 10 15 20 MEAN WATER TEMPERATURE DIFFERENCE.21 H = heat loss through the exposed edge. The rate of heat loss from the surface of the inner slab is proportional to the temperature difference between the floor surface temperature and air temperature. t„ = outdoor design temperature. L = exposed edge. 5. Panel performance is directly related to room conditions.4x2.4x3. Calculate room heat loss.4x]. W. 7. ts = floor temperature. Determine room design dry-bulb temperature. A Using electric slab heating as a primary system.tS m-'-C/W-lf f M B t » X 0 . basements and family rooms.inner floor area = total floor area minus floor area included within 3-ft (0. the heat output of the inner floor equals the heat input minus the downward heat loss. The balance of the design can be determined from the data in Chapter 11. W/ft>F(W/m2•°C). For slab tempering in kindergartens. using the slab surface temperature found in Step 3 and the factors in Chapter 23 of the 1981 A » / *$r j. Determine panel area for heating. U = heat transmission coefficient. 8. The rate of heat loss is also affected by the floor material (concrete. panel type. will dissipate into the area to be heated by radiation and convection heat transfer.6 W/ft 2 (6.2 and 3. 10. bathroom floors. Determine water flow rate and pressure drop. •Derived from Fig. poorly insulated buildings (farrowing.9-m) border along exposed edge. ^f = h0 + 11 = 2.Panel Heating and Cooling Systems 8.BtttA) • H* X 3. Calculate room sensible and latent heat gains. PANEL COOLING SYSTEM DESIGN The heat loss through the inner slab at a distance beyond 3 ft (0. water flow rate and panel arrangement. The procedure is: 1. Assuming a steady-state condition. ta = air temperature. the total heat loss through a slab on grade can be divided into two segments: (1) the heat loss through the exposed edge including the floor area within 3 ft (0. ft (ro) Sensible Gain. Assuming the heat loss for the inner floor to be 1 W/ft 2 (11 W/m 2 ): h. tf = W/ft. W/ft2 (W/m2).4> 11x12x8(3. If room db and rh are allowed to drift for energy conservation purposes.) + 11 Design Steps for Metal Ceiling Panels Panel design requires specifications of panel area.4 W/m 2 ). Calculate latent cooling available from the air.-t0)+\A where H.5 (/.=KL(t.7x2. ft2 (m2). supply water temperature.17. i .dewpoint of 55 F (12. The room heat gains are: A B C Dimensions. (19) H.= indoor design temperature. Thus. Select mean water temperature for cooling. W/ft2 (W/m2). F (°Q.6x6./. Calculate sensible cooling available from the air. Blu/h(W) Latent Cain. Previous work on the subject. the highest expected dewpoint should be used.000(4 102) 400(117) 500(147) 1. A. 9 30 \X ^ FUNDAMENTALS VOLUME.9 m) of the exposed edge is independent of outdoor temperature. formance curves can be used. /. . Step I: Design Conditions Referring to Example 1 for design procedures for a plaster ceiling heating panel system. Determine panel cooling load.= h0 + 1 = 0. brooding and milking parlors) an input of 30 to 40 W/ft 2 (323 to 431 W/m 2 ) is a normal design range. F (°C). Panel Cooling Performance* D-22 . Air-side design also must be established. W/ft2 (W/m2). 9. F(°Q. A . For spot heating in unheated.8 War . resulting in a.F<\V/m-°C).84 (/s . are each to be maintained at 78 db and 45% rh.18 indicates that heat losses through the inner floor [area beyond 3 ft (0. Design the panel arrangement. a design range of 10 to 15 W/ft 2 (108 to 161 W/m 2 ) of slab area is the maximum input recommended.8°C). 10.9 m) of the edge] remain nearly constant at approximately 0.. Step 2: Heat Gains Calculate the room heat gains by the methods outlined in the 1981 FUNDAMENTALS VOLUME. Btu/h (W) Ilx|2x8(3. 11.000(293) s (21) vV (21a) Electric slab heating systems can also be designed by using part of the procedure for hot water panels as a guide. Follow the procedure for hot water panels. relative humidity and dewpoint. 4. = total slab loss. a range of 1 to 5 W/ft 2 (11 to 54 W/m 2 ) is recommended. Select mean water temperature for heating. Determine panel area for cooling. h. Step 4 Slab Heat Loss Determine the heat loss from the slab.) + 1 In SI. F(°C).4) 3. Steps 1.-. 11 (0.7) 24.43 g/kg) of dry air.071 (0.760(516) 4.1m2) Typer = 0.0) 71(6.81 = 46 ft2 (4.58.i) If panel type r = 0.344 = 6.180(1 225) 18.0 Btu/h-ft2 (56. Based on the available ceiling area.344(2 148) 10.760 V B (i u 516) /hfW)/ 4.071 (0.4) 106(28.31 (0.6) 34. K panel type r = 0. Blu/h.740 Btu/h (803 W) Panel Cooling Available.1) 112(12.1) Available Panel Area.0)1.286 Btu/h (670 W) Panel Cooling Required from typer = 0. Assuming panel type r = 0.S8.1) Panel Area Available. The room temperature minus the mean water temperature is 78 .0( 75. the cooling available satisfies the panel load and no increase in air flow rate is necessary. the moisture pickup of the air is sufficient to offset the room latent gain.8) 127(11.1) 97(9.012) is used. ft2(m2) X X X Room Panel Performance.107) 0.31 (0. and the panel area is reduced to: 2.0 gr/lb (1.h/Btu (m 2 -'C/W) 0.11 (0.31 (0.4) 106(28.31 (0. The calculations are: Panel Cooling Load = 2.1 (78.11 (0.0)0.055)1 exceeds the cooling required.055) plus 127 .071 (0.500(1 119) Blu/hfW) / 14.107) 0.68 544.0-58.0 ( 56.760(516) 1.72 g/kg) of dry air.4) 544.8°Q.0 .61 (0.0(107. Blu/h(W) (78.54.055).61 (0.118(1 265) 127(11.1) 112(12.54.8.61 (0.107)] is deficient in capacity and panel type lr .61 (0.0(178.8) 1 ( 0.500(1 026) I Heat Gain. the panel area for Room B consists of 81 ft2 (7. r = 0.8°Q temperature rise .107). Btu/h (W) 112(12.0 . the panel cooling available for Room A is: Room sensible load Panel cooling Air cooling required Revised airflowrate = 14.740(510) Pand Area. from typer = 0.7) 24.0(159.107) or r = 0.0 and 34.8) 5114. Heat losses are the same. The panel cooling performance at 19.7) 24. ft 2 (m 2 ) I8.2861 670) 1. Assume a 5 deg F (2.6) 14.5 F (U. 4.0(159.0)0.61 (0.0.8) 80(38) Similarly for Room B: 190(90) Step 5: Latent Cooling Assume a central station air system supplying conditioned air to the rooms at 58 F (14. Biu/h(W) (64.h-F/Btu (m 2 .61 (0.0)1. Fig.6) 14.8) 127(11.1) 2.'C/W) 0. cfro(LA) gr/lb (g/kg) air Supply Air Latent Cooling. 24.6 m2) The data for Room C are: Step 7: Panel Load The panel load is the room sensible load minus the cooling done by the air.61(0.000(4 102) - XsupplyAir\ I Sensible 1 1.012) Panel Cooling Load.3 m2) of panel type r = 0. If either of these two ceiling types is still desired for architectural reasons.8) ROOMC Panel Resistance. Btu/h(W) 1. = 56(13.7) giving a mean water temperature.0. adjustments in supply air quantity or design temperatures would have to be made accordingly. Btu/h (W) 2.055) 0.22 CHAPTER 8 1984 Systems Handbook Step 3: Select Mean Water Temperature F(°Q = 55(12.012) \ Sensible 1 _ 3.h/Btu (m 2 -"C/W) 0.740(510) 1.4) Panel Cooling Available.31 (0.3) Room dewpoint temperature.0 gr/lb (7. The latent load capacity ofthe'air is: Air Flow Rale.6) ( ' Room \ _ Humidity I 80(18) 80(18) 190(90) /" Supply Air \ I Humidity I In all three rooms.820(2 877) Step 8: Panel Area The panel area is a function of the mean water temperature and the type of panel used. Step 6: Sensible Coolingfrom Air ( Air Flow Rate cfm (L/s) 80(18) 80(18) 190(90) FfO Supply Air Sensible Cooling.5 deg F (10. 2 ft .9°C) wb.000 Btu/h (4102 W) = .404(1 045) The panel cooling available from either panel type.0 ( 75. ft 2 (in 2 1 « " « Pand Cooling Available.fc 2 (W/m 2 ) Panel type [r = 0.0-58. Inlet water temperature. additional cooling must be provided by increasing the supply air quantity.0 ( 75.5 m2) of panel type r = 0.055) does not satisfy the panel cooling load for Room C in Step 7.5 (14.0(107.0(107.055) (24-18) Thus.5) / .055) 0.5 deg F(I0.176 ( 697) 1.8°Q temperature difference for the three types of panels is 18.740 ( 801) 9.0 .071 (0.6) 14.740 ( 510) 2.61 (0.68 (64. Btu/fc(W) Pand Cooling Performance.168 ( 910) 4. D-23 .055) 0.31 (0.0 ( 75. Btu/h.656 Btu/h (1950 W) = 6 ' 6 5 6 = 308 cfm (145 L/s) (20xl.0 ( 56.0) I.292.0)0.ft 2 (W/m 2 ) All three types of panels exceed the panel cooling load. = 58. cfm (L/s) 80(38) A B C Room Pand Resistance 2 rt >F-h/Btu (m 3 -*C/W) 106(28.012) Panel Performance.508(1 612) 7. Minimum supply air quantities are: Room 0.0( 56.772/34.012) is used.107) = -2. and thereby a moisture difference between room and supply air conditions of 10.F. ( Pand Performance.071 (0.0-54. respectively.7.68 (64.5 = 19.1 W/m2).m2) 1. resulting in a moisture content of 54.4) 1. Step 9: Room Heat Loss See Step I for plaster ceilings.61 (0.0 ( 56. a combination panel may be used.0(107. 75^6 and 107.107) 0. 2 f! .180(1225) Room \ Temp I / Supply Air I Temp \ I ROOMB Panel Resistance.1 1. 27 shows design panel performance for three types of ceiling panel systems.048| 891) 4.071 (0. Btu-hlW) 5.740(510) 1.760( 516) I Cooling. Since these panels have the same appearance and construction. I 1. reducing the design panel area as follows: Minimum Supply Air Quantity.7.055) 0.0 = 82 ft2 (7.107) 0.055) = 454 Btu/h (133 W) = 4S4 Area of Panel ^ ^ 76 ft2 (7. If this were not the case. 2 fl -F. Btu/h-ft 2 (W-m2) 18. Bm/h-ft2(W.488(1 117) ROOM A Pand Resistance.1) Pand Coating Available.I (78.71 24.012) Available Pand AreaL 2 f t '(m ' 2> 18. Step 4: Minimum Supply A ir Quantity Minimum supply air quantities should be based on the recommended practices as given in the ASHRAE HANDBOOK and as dictated by local codes. the calculation is: Pand Cooling Load.4°Q db and 53. the cooling available exceeds the panel cooling load. p. 189). P. Leopold: The mechanism of heat transfer panel cooling heat storage—Part II. Parmelee and J. 61. 144).1952. L. 73. R. 58. L. Jr. 74.C.S. 57. Spiegel: A water-cooled luminaire in a panel-air system (ASHAE Transactions. and T. D-25 . June 1966.F. Schutrum: The ASHVE environment laboratory (ASHVE Research Report No. Vol.P. p. 123). Vol. C M .: The effect of asymmetric radiation on the thermal and comfort sensations of sedentary subjects (ASHRAE Transactions.F.F. Jr.. 139). Schutrum and T. 1444. 62. Rapp. May 1967. T. 95).1951. Vol. p. p. Leopold: The mechanism of heat transfer panel cooling heat storage (Refrigerating Engineering. 1967. p. 63.Conditioning. McNall.D. June 1948. 351). G. and R.C. Min. p. Schutrum and T. Humphreys.F. Gagge. p. Min:. Biddison: Thermal and comfort sensations of sedentary persons exposed to asymmetric radiant fields (ASHRAE Transactions. 109). A. W. Parmelee.F. 33).24 CHAPTER 8 C. McNall. and L. John Vouris. p. G. G. Lighting and cooled air effects on panel cooling (ASHAE Transactions.V. p. J . Schutrum.1).1970. ASHVE Transactions. p.F.S. 111.4.D.1955. Vol. Vol.1958.V. Vol. p.E. Vouris: Natural convection and radiation in a panel-heated room (ASHAE Transactions. Fanger: Calculation of thermal comfort: Introduction of a basic comfort equation (ASHRAE Transactions.E.C. July 1947. C. Leopold: Design factors in panel and air cooling systems (ASHAE Transactions. and J.E. Vol. Part 11.H. 76. Schlegel and P. p. 64. Min: Preliminary studies of heat removal by a cooled ceiling panel (ASHAE Transactions. Vol.S. Solar radiation (Refrigerating Engineering. 1984 Systems Handbook L. 571). p.C. Schutrum. L. Cyril Taster. 63). 337). C.O.1968. Boyar: The influence of radiant energy transfer on human comfort (Heating. 61). Min: Cold wall effects in a ceiling-panelheated room (ASHVE Transactions. Vol. P. Vol.1956.1957.C. Piping & Air.1958.E. Hardy: The effective radiant field and operative temperature necessary for comfort with radiant heating (ASHRAE Journal. 187). 64.8. p. 62.V.3) 27 (2 J ) 67(6. Fanger: Thermal Comfort Analysis and Applications in Environmental Engineering (McGraw Hill. From Fig.O. two panels are designed. Parmelee.S.F.P. Humphreys: ASHVE Research Report No.T. and C M .F. © . 285). p. G. 6 G. Schutrum.L. Schutrum and CM.fC) Btu/n. Leopold: Hydraulic analogue for the solution of problems of thermal storage. Vol. Gilkey.3°Q for ceiling type r = 0. Vol. Vol. 389). For example.. For Room B. p. Vol. R. 69. 10 L. 28. and the panel area within 4 ft (1. Sartain and W. 4 for the Design curve. p. 121). 1600—Thermal design of warm water concrete floor panels (ASHAE Transactions. Humphreys: ASHAE Research Report No. "H. W. 63. ft 2 (m 2 > 98(9. p. This should be done in cooperation with the architect and manufacturer.1947. P. 59. 337).L. 37). 455). Chairman.107) 0.H. Vol. 1576-Natural convection and radiation in a panel heated room (ASHAE Transactions. Parmelee and R. Thermal Comfort Conditions (ASH RAE. Btu/h <W) 6. Chairman. Schutrum and CM.1942. »L. 14 Standard Specifications for Gypsum Plastering.1956. p. air outlets.V. a mean water temperature of 164 F (73.300(1850) 2. P. 4 T. Hutchinson: Influence of gaseous radiation in panel heating (ASHVE Transactions. A water temperature is selected to give a panel output of approximately (6300 x 0.6 Btu/h-ft2 [(1846 x 0. Maher. Gordon.1956. 28 Metal Ceiling Panel Design Graph Showing Panel Surface Temperature and Mean Water Temperature vs.F. Humphreys: ASHVE Research Report No. 62. Vol.62. R. Alberty: Temperature and Heat-Loss Characteristics of Concrete Floors Laid on the Ground (Research Report'48-1.00012300) * * v 71 (224) 90(284) 119(375) Panel Area.D.C.C Houghten et al: Heat loss through basement floors and walls (ASHVE Transactions.1.055) 0. L. „ . 23 for typical split-panel arrangements. Boyar: Room temperature dynamics of radiant ceilings and air conditioning comfort systems (ASHRAE Transactions.F. Parmelee.5/4) = 2300 W/m 2 ). Room A B C Cooling Panel Area. The panel pressure drop must be obtained from manufacturer's data.2) Step 12: Hydraulic Data The water flow rate may be based on a 20 deg F (1 l. Parmelee. Vol. 13 R.E. University of Illinois). Schutrum and J.61(0. „ .1957. p. As can be seen from Steps 1 Standard 2 55-74.107) 0. Schutrum.G. A 42. Room C is similar to Room B.3 = 4 m 2 ). Fig. based on 70 F and 70 AUST. A. Harris: Performance of covered hot water floor panels.23 8 and 11. 245). Macey: Heat loss through a solid floor (Institute of Fuel Journal. 1490—Heat exchanges in a floor panel heated room (ASHVE Transactions.E.3) 27(2. Maher.F. 495). G. and J. Min et al: ASHAE Research Report No. 1516—Effects of room size and non-uniformity of panel temperature on9panel performance (ASHVE Transactions.Panel Heating and Cooling Systems 8. Van Nieukerken. and by ASHAE Laboratory Staff Members. Step 11: Heating Panel Area The heating panel area for each room is given below. Vol. The temperature may be increased using experience and judgment if desired. Vouris: ASHVE Research Report No.2 m) of the outside wall is 4 x n = 44 ft2 (1. W. Huebscher: Forced convection heal transfer from flat surfaces (ASHVE Transactions. 1974). p.500 ( 730) 8.012) For Room A. Vol. which is similar to Fig. This represents a suggested minimum mean water temperature for comfort.1954. 369). Room A B C Panel Performance. Inc.1 6 4 F(73. '*A Subcommittee of the TAC on Panel Heating and Cooling.31 (0.B.F. BIBLIOGRAPHY C. a mean water temperature of 164 F (73. the areas are close enough so that the larger area is used as a common panel for heating and cooling.2) Ceiling Type r 0. By inspecting the data for rooms B and C and Fig. Gilkey. (102 mm) on center. 1947. E.1953. 1499—Effects of non-uniformity and furnishings on panel heating performance (ASHVE Transactions.B.1956.l°C) temperature difference between inlet and outlet temperatures or any temperature difference suitable to the design of the piping distribution system. Small Homes Council. p. 8 except the performance is based on the convection output from Fig. D-24 . p. sprinkler and smoke devices.1946). assume Room A ceiling to be aluminum panels with pipes 12 in. The suggested basis for selecting the mean water temperature is that the area of panel adjacent to and within approximately 2 to 3 ft (0. and J. 1559—Thermal design of warm water ceiling panels (ASHAE Transactions. 60.300 Btu/h (1846 W). Peterson: Radiation and convection from surfaces in various post ions (ASHVE Transactions. and one panel for cooling only = (81 . one panel for heating and cooling = 27 ft2 (2.fl2(W/m2) Heai Loss.5/44) = 71. Room B ceiling to be aluminum panels with pipes 6 in. Fleming. and by ASHAE Laboratory Staff Members. Part 1-Thermal characteristics (ASHAE Transactions. p. 1473—Heat exchanges in a ceiling panel heated room (ASHVE Transactions. E. Gordon.F. Schutrum. p. 1963. and Room C ceiling to be aluminum sheet with copper tubes4 in. 12 E.B. Chapman. New York. 13 A Subcommittee of the TAC on Panel Heating and Cooling. 59. P.M.3°C) is selected. 22-128. Step 13: Panel Arrangement The panel and piping arrangement must be designed to accommodate the various elements in the ceiling such as lights. Vol. Snyder. Van Nieukerken.M.M. W. Humphreys: ASHAE Research Report No. 369). Room A transmission loss equals the heat loss of 6.0) 89(8. p. Vol. 44.2 m 2 ). NY. 513). 53.071 (0.6 to 0. See Fig. H.61 (0.2 x 3. and CM. and B. p.F. 71). speakers. 54. 197).F.3) 81 (7 J ) 291 (27. 48.5 m 2 ). Bareither. 7 L. Chapman. radiation. (305 mm) on center. 28. Vol. 1972). Snyder.V. REFERENCES t'wrtoe) "I I 100 I I 120 I 1 I 140 T P I 160 0R T 1 180 1 200 M»T * Derived from Figs.L. Vol. 5 G. "H. L. convection and conduction (ASHVE Transactions. 2 and 4. the panel area for heating may or may not be the same for cooling.T. p. Vol.27) + 46 = 100 ft2 (9.P.1954.S. and C M . Wilkes and C. (152 mm) on center. 1938.D. p.61 (0. I8 F.N. Humphreys: ASHVE Research Report No.55). 60.11) will satisfy the requirement. ft2(ra2) 89(8. Panel Heating Performance* Step 10: Mean Water Temperature Typical design panel performance at various mean water temperatures is given in Fig.5) 67(6. 3 F.V.1) 46(4. Including Requirements for Lathing and Plastering (American Standards Association.1948. ft 2 (m J ) Healing Panel Area.1953. H. G. 53. 239). L. 28. For purposes of simplification.9 m) of the outside wall should have an output of not less than 50% of the wall transmission heat loss. 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